Description

3
rd
World Conference on
Applied Sciences, Engineering & Technology
27-29 September 2014, Kathmandu, Nepal
WCSET 2014067 © BASHA RESEARCH CENTRE. All rights reserved.
http://basharesearch.com/wcset2014.htm
Waste Heat Recovery System for a Heavy Duty Six Cylinder Engine
CHANDRASHEKHAR BHAT
1
, SHARMA S. S.
1
, JAGANNATHA K.
1
, ACHUTHA KINI
1
,
SUNIL KUMAR PANDEY
2
, SARAVANA VENKATESH
2
, PRAVEEN KUMAR K S
1
1
Manipal Institute of Technology, Manipal, Udupi, Karnataka, Ind

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3
rd
World Conference on Applied Sciences, Engineering & Technology 27-29 September 2014, Kathmandu, Nepal
WCSET 2014067 © BASHA RESEARCH CENTRE. All rights reserved. http://basharesearch.com/wcset2014.htm
Waste Heat Recovery System for a Heavy Duty Six Cylinder Engine
C
HANDRASHEKHAR
B
HAT
1
,
S
HARMA
S.
S.
1
,
J
AGANNATHA
K.
1
,
A
CHUTHA
K
INI
1
,
S
UNIL
K
UMAR
P
ANDEY
2
,
S
ARAVANA
V
ENKATESH
2
,
P
RAVEEN
K
UMAR
K
S
1
1
Manipal Institute of Technology, Manipal, Udupi, Karnataka, India
2
Ashok Leyland Technical Center, Chennai, India E-mail: chandra.bhat@manipal.edu, ss.sharma@manipal.edu, jagan.korody@manipal.edu, achutha.kini@manipal.edu, Sunilkumar.Pandey@ashokleyland.com, Saravanavenkatesh.r@ashokleyland.com, kpraveenkmr22@gmail.com
Abstract:
Organic Rankine cycle (ORC) is seen as an economical and environmental friendly technique to waste heat recovery (WHR) from the exhaust gases. A mathematical model of the ORC system is developed in MATLAB Simulink and is then coupled to the engine model in AVL boost which simulates the working of the engine and the ORC and leads to the computation of BSFC. Using the Simulation results, the influence of working fluid flow rates, the engine speed and the injection timing are studied. The simulation results show an improvement in brake specific fuel consumption (BSFC) to an extent of 4.5% (equivalent to 11 kW ORC power) when the working fluid flow rate is 0.65 kg/sec and at a lower working range of 1900 to 2200 engine speeds in RPM. When the working fluid flow rate is 0.4 kg/sec a 3% improvement in BSFC is observed at a greater working speed. For the ORC model to work for a higher range of engine speeds from 1300 to 2200 RPM, the working fluid flow rate of 0.4 kg/sec is taken to be the optimised value thereby improving the BSFC by 3%.
Keywords:
Organic, Cycle, Design, Heat, Simulation.
I
ntroduction:
DIESEL engines are the internal combustion engines, which require high power density to improve the overall efficiency. Diesel engines remove significant amount of combusted fuel energy through the exhaust of the engines. Out of the total waste heat generated in these engines, 40 % is wasted as exhaust gases [1] with a maximum temperature of 650° Celsius [2] depending upon the type of diesel engine and its working conditions. Great strides are being made in utilizing this fuel energy that is wasted through the engine exhaust gases using different waste heat recovery technologies. They also bring added benefits of reduction in emissions and CO
2
footprints from the exhaust gases. There has not been any significant improvement in the past few decades when it comes to waste heat recovery (WHR) techniques in truck diesel engines. The challenge lies in the fact that the exhaust gas heat source varies continuously from low to high as the engine speed varies. Hence, design and developing a WHR system for such an engine is necessary. By developing a WHR system, the overall power (Power of Engine + Power of WHR) of the engine can be increased and the BSFC of the engine be reduced that can bring economic benefits to the vehicle industry and the customers. The heavy-duty diesel engines are the most widely utilised engines. A heat balance pie chart (Fig. 1) for a heavy-duty six-cylinder diesel engine shows nearly 22% of the fuel energy being wasted as exhaust gas heat. In this engine, the exhaust temperature varies between 400 - 600°C and exhaust mass flow rate between 0.18 - 0.4 kg/sec from lower to higher engine speeds. It is necessary to utilise this waste heat to produce useful power, which may drive some engine accessories. This may be achieved by identifying a better WHR technique and a working fluid for this engine.
Figure. 1
Heat balance Pie Chart of the six cylinder Diesel engine
The investigation involves (a) Identifying a suitable WHR technique and a corresponding working fluid having high heat transfer and environment friendly characteristics, (b) Development of WHR model using Matlab Simulink and then coupling it with the engine model developed in AVL BOOST and finally simulating both the models simultaneously to compute the brake specific fuel consumption (BSFC) of the engine, and (c) study of influence of working fluid flow rates, the engine speeds and the injection timings on the BSFC improvement of the engine, using the simulation results. Prior research indicates that,
CHANDRASHEKHAR BHAT, SHARMA S. S., JAGANNATHA K., ACHUTHA KINI, SUNIL KUMAR PANDEY, SARAVANA VENKATESH, PRAVEEN KUMAR K. S. Proceedings of the 3
rd
World Conference on Applied Sciences, Engineering and Technology 27-29 September 2014, Kathmandu, Nepal, ISBN 13: 978-81-930222-0-7, pp 371-377 Pannu, et al
[3] studied the different WHR techniques that could be utilized to recover the waste heat of the diesel engines and gas engines and concluded that ORC system is a better technique to tap the waste heat due to its simplicity and environment friendly
characteristic’s. In addition, its
low-pressure range makes it economically beneficial and simple to construct with only one working fluid. Research Prior to that [4, 5, 6], have reported that use of ORC has indeed leads to good recovery of heat from the waste gases. Mago et al. [7] stated that dry and isentropic fluids such as pentane and R245fa have a better thermal efficiency when compared to other fluids when used in ORC cycle. Wang et al. [8] have stated that R245fa as a working fluid for ORC will be beneficial both in terms of heat transfer and environment friendly characteristics. G. Kosmadakis [9] have concluded that refrigerant R245fa has a better safety level and has a longer atmospheric lifetime along with a lower (ozone depletion potential (ODP). The prior research indicates that ORC can be selected to be the best environmental friendly waste heat recovery technique in comparison to the other methods. The working fluid R245fa is an environmental friendly, isentropic and less toxic as compared to other fluids, hence is seen as the best ORC working fluid. The ability of ORC system to adjust to varying temperature heat sources also makes it suitable to be used in the truck engines.
Organic Rankine Cycle:
An ORC (Fig 2) consists of an evaporator, turbine, condenser, pump, and a working fluid. In this, the evaporator or vaporizer acts as a heat exchanger and facilitates the working fluid to evaporate using the external heat source. The vaporized working fluid passes through the turbine where work is produced by the expansion of vapor. This vapor is then condensed to the saturated liquid state and finally pumped back to the evaporator.
Figure 2
Block Diagram of ORC
Modeling and Simulation of WHR System:
Functioning of each component of the WHR system is mathematically modeled by making few assumptions. Evaporator is modeled as a shell and tube heat exchanger (counter flow) as shown in Fig 3 to determine its effectiveness and also to check whether the condition of working fluid at outlet of the evaporator is in vapor form.
Figure 3
Longitudinal section of Evaporator
Where t
hi
, be the temperature at the inlet of the shell side fluid. t
ho
, be the temperature at the outlet of the shell side fluid. m
w
, be the mass flow rate of the tube side fluid. t
wi
, be the temperature at the inlet of the tube side fluid. t
wo
, be the temperature at the outlet of the tube side fluid. ------------------------- [1] Where N
uin
is Nusselt number, K
in
is thermal conductivity and d is the diameter of tubes. Shell side heat transfer coefficient is given by -----------------------[2] Where, K
out
is thermal conductivity and d
h
is the hydraulic diameter of the shell. Overall heat transfer coefficient is given by ------------------------------[3] Effectiveness of the heat exchanger is computed from effectiveness NTU method. Effectiveness for counter flow condition is given as: ------------------[4] For parallel flow, effectiveness coefficient is given as ------------------[5] where Cr = C
h
/C
c
and NTU = (A
o
+ U
o
)/C
min
A
o
is the overall area of heat transfer and C
min
is the minimum heat capacity rate of the shell side and tube side working fluid, C
h
=m
h
*C
ph
= heat capacity rates for the shell side fluid, C
w
= m
w
*C
pw
=
heat capacity rates for the tube side fluid. Obtain value of enthalpy of working fluid at outlet of evaporator (
g
3
) using the effectiveness, and energy balance equation:
Waste Heat Recovery System for a Heavy Duty Six Cylinder Engine Proceedings of the 3
rd
World Conference on Applied Sciences, Engineering and Technology 27-29 September 2014, Kathmandu, Nepal, ISBN 13: 978-81-930222-0-7, pp 371-377 ------- [6] Where m
h
is the mass flow rate of the shell side fluid. Also effectiveness e is given by ----[7] Where m
w
, is the mass flow rate of the tube side fluid. If the value of g
3
corresponds to superheated state, then mass flow rate of working fluid may be accepted. Generally vary the mass flow rate of the working fluid till desired superheated state of the working fluid is achieved. The turbine is modeled as a radial inlet and axial outlet turbine. The actual turbine work is obtained based on the empirical relations as follows, ------------[8] Where
g
3
is the enthalpy at turbine inlet.
g
4
is the enthalpy at turbine outlet.
turb
is the efficiency of the turbine.
turb
=
W
actual turbine /
W
isentropic turbine ------[9] Where,
W
actual turbine
is the actual work that the turbine develops.
W
isentropic turbine
is the maximum possible isentropic work. Fig. 4 shows the condenser which is modeled as shell and tube heat exchanger [10] and the outlet temperature of the cooling fluid is fixed. Mass flow rate of the cooling fluid is varied to achieve the desired cooling effect that takes place at constant pressure. The process involves calculating the overall heat transfer coefficient and effectiveness for both the de-superheating and condensing region and the methodology of calculating the condenser effectiveness is almost similar to that of heat exchanger.
Figure. 4
Condenser Model
The tube side condensation heat transfer coefficient is given by Chato correlation [11] for -[10] Where
1
, is the density of working fluid at saturated liquid condition.
v
, is the density of working fluid at saturated vapor condition,
1
, is the dynamic viscosity of working fluid at saturated liquid condition, T
sat
, is the corresponding saturated temperature, d
h
, is the hydraulic, diameter, h
iv
, is the modified latent heat and is given as, ---------------- [11] For Re > 35000, Akers, Deans & Crosser propose the empirical correlation for average condensation heat transfer coefficient on the inside of the horizontal tube of diameter D; -------------------------- [12] ------------------ [13] Where
M
l
, is the mass flow rate of liquid in kilograms per second.
M
v
, is the mass flow rate of vapor in kilograms per second.
R
el
, is the Reynolds number considering liquid flow.
R
ev
, is the Reynolds number considering vapor flow.
l
, is the dynamic viscosity at saturated liquid condition.
v
, is the dynamic viscosity at saturated vapor condition Here m
r
=
M
l
,
=
M
v
as the flow rate is assumed to constant and system is a closed loop system. The shell side heat transfer coefficient for the de superheating region and condensing region remains same as the flow is single phase flow as the basic assumption used here is that the coolant medium i.e. water is not allowed to reach the vapor state (limited to 343 K). The equation 2 will be used for the calculation of heat transfer coefficient for the shell side in both the de superheating and condensing region. To be specific, condenser effectiveness followed by cooling fluid inlet temperature (T
v2
) is obtained by same approach as motioned above (1 to 5). Then for the condensing region calculate the heat transfer coefficient for the tube side based on the equations from 10 to 13. Then using T
v2
as the outlet temperature of cooling fluid at the condensing region obtain the heat transfer coefficient for the shell side based on equations from 2 to 3. Then obtain the effectiveness for the condensing region based on the equations 4, 5 and 7. With this the cooling fluid inlet
CHANDRASHEKHAR BHAT, SHARMA S. S., JAGANNATHA K., ACHUTHA KINI, SUNIL KUMAR PANDEY, SARAVANA VENKATESH, PRAVEEN KUMAR K. S. Proceedings of the 3
rd
World Conference on Applied Sciences, Engineering and Technology 27-29 September 2014, Kathmandu, Nepal, ISBN 13: 978-81-930222-0-7, pp 371-377 temperature T
v1
at the condensing region is obtained using equation 6. The above steps are to repeated by varying the mass flow rate in order to obtain a cooling fluid temperature T
v1
equal to its ambient temperature liquid state i.e. 303K.
Simulation:
The model of the simulated Six Cylinder Diesel Engine developed using AVL Boost as shown in Fig 5.
Figure. 5
AVL Boost Model of Six-Cylinder Diesel Engine
The AVL Boost engine model is composed of several pipes (1-29), intake manifold (PL1), 6 combustion cylinders (C1-C6), junctions (J1-J7), air filters (CL1), Charge air cooler (CO1), measuring points (MP1-MP26), system boundaries (SB1-SB2), exhaust gas recirculation cooler (CO2), turbocharger (TC1), selective catalytic converter (CAT1), muffler (PL2). The air from atmosphere is sucked into the compressor of the turbocharger via the air filter, the air gets compressed and its temperature increases then passes through the Charger air cooler where it gets cooled by the atmosphere air and gets increased in density and moves into the inlet manifold PL1 where it passes into the individual cylinders and gets combusted along with fuel to produce the work on the crankshaft. The exhaust gases that are generated during combustion pass into the turbocharger and finally go into the atmosphere through the catalytic converter and muffler. Since the exhaust gases coming through the outlet of turbocharger are at a temperature of 400 to 600 degrees Celsius. This high amount of waste heat goes into the atmosphere, which can be utilized The ORC waste heat recovery system for the corresponding engine is developed using MATLAB SIMULINK [12] and the two models were linked together using DLL MATLAB [13] function available in AVL Boost. The six-cylinder diesel engine model is developed in AVL BOOST 2011 [14]. Fig 6 shows Simulink model of ORC.
Figure. 6
Simulink Model of ORC
The actual thermodynamic modeling for pump and turbine are not taken into consideration during the simulation. The turbine is assumed to be 75 % efficient with no flow losses. The enthalpy after turbine is directly obtained using the look up table module. For the simulation it is assumed that pump work consumed is 0.5 kW. The turbine power is obtained based on the evaporator pressure in the system i.e. 5, 10, 15, 20 bar as in Fig. 6. The condenser is simulated with water as the cooling medium. The simulation model of evaporator is based on the dimensions of a standard heat exchanger dimensions used in one of the engines. The evaporator model optimized by studying the influence of working fluid flow rate, exhaust temperature and exhaust mass flow rate on the working fluid enthalpy at outlet. The inputs for the evaporator model are exhaust temperature and exhaust mass flow. The refrigerant mass flow is the user input in Matlab Simulink Block of Evaporator before any run and the result is obtained. There is no pressure drop assumed in the system. With this, the overall heat transfer coefficient is obtained for the evaporator system. In order to know the effectiveness of the evaporator, the effectiveness NTU relation for counter flow is used. Finally the enthalpy after evaporator is obtained as the output. In the subsystem model of turbine, the enthalpy after evaporator is used to calculate the turbine work assuming the turbine to be 75 % efficient. The model is based on the maximum pressure (i.e. evaporator pressure) of ORC. If the maximum pressure is 5 bar then the enthalpy after evaporator should be greater than 450 kJ/kg (to have a saturated vapor at turbine
inlet), else the block won’t proceed thus producing no
output. Similarly for maximum pressure of 10 bar, 15 bar and 20 bar the corresponding enthalpy after evaporator should be greater than 470 kJ/kg, 480 kJ/kg and 488 kJ/kg for the block to run and produce the desired turbine work output. After this stage the enthalpy after turbine expansion is transferred to the condenser model
.
In the condenser mode, two inputs are needed the enthalpy after turbine and refrigerant mass flow rate. The refrigerant mass flow rate is obtained from evaporator model and enthalpy from turbine model. The values move into the individual blocks of condenser for ORC maximum pressure of 5 bar, 10

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