On the effects of increased coolant temperatures of light duty engines on waste heat recovery

Applied Thermal Engineering 172 (2020) 115157

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On the effects of increased coolant temperatures of light duty engines on waste heat recovery

T

Vikram Singha, , Jelmer Johannes Rijpkemab, Karin Munchb, Sven B. Anderssonb, Sebastian Verhelsta ⁎

a b

Faculty of Engineering, Division of Combustion Engines, Lund University, 21100 Lund, Sweden Chalmers University of Technology, Hörsalsvägen 7B, 41280 Göteborg, Sweden

HIGHLIGHTS

in system efficiency using waste heat recovery up to 5.2 percentage points. • Increase temperature increase shows up to 1.7 percentage points gain in efficiency. • Coolant coolant temperature dependent on engine load and speed was observed. • Optimum • Best working fluid for coolant heat recovery simulated was cyclopentane. ARTICLE INFO

ABSTRACT

Keywords: Low temperature waste heat recovery Elevated coolant temperatures Light duty engine Rankine cycle Recoverable power Reduced heat losses

In this paper, an investigation is done into the potential of increasing the coolant temperature of an engine to maximize the powertrain efficiency. The study takes a holistic approach by trying to optimise the combined engine and waste heat recovery system. The work was done experimentally on a Volvo 4-cylinder light duty diesel engine in combination with Rankine cycle simulations. For the study, the coolant temperature was swept from 80 °C to 160 °C at different operating points. It was seen that with increased coolant temperatures, the brake efficiency of the engine increased by up to 1 percentage point due to reduced heat losses. An optimum coolant temperature was observed, dependent on the operating point, for maximizing coolant recoverable power. An expansive study was done simulating 48 working fluids for a dual loop waste heat recovery system. From the working fluids simulated, cyclopentane was seen as the best for coolant waste heat recovery, whereas methanol and acetone were better for the exhaust gases. The gain in efficiency seen, was up to 5.2 percentage points, with up to 1.7 percentage points as the effect due to recovered power from the coolant.

1. Introduction Over the past few decades, the internal combustion engine has moved towards higher fuel efficiencies to reflect the trends in the market as well as changes in emission regulations. While engines today are quite fuel efficient with heavy duty (HD) diesel engines having indicated efficiencies of 50% or higher [1], a large part of the fuel energy is still wasted in the form of heat – from the exhaust, coolant, oil and radiative losses to the environment. These losses can be reduced by optimising the combustion chamber, the spray and the gas exchange processes and the combustion mode [2–4]. However, these losses still form a large fraction of the fuel energy that is not transmitted to the engine crankshaft. One method to increase the power output at the crankshaft is to use a secondary thermodynamic cycle to recover some of the waste thermal ⁎

energy from the engine. Thermodynamic cycle based waste heat recovery devices have been researched quite well and have shown good potential for implementation [5]. It should be noted that waste heat recovery has been used in other applications as well, such as to reduce warm up time for the engine oil [6,7]. However, in these cases, it does not affect the steady state power output at the engine crankshaft. A waste heat recovery system, such as the Rankine cycle, can be used to take heat from any of the engine’s heat sources (intercooler air, oil, water or exhaust) to heat a secondary fluid at higher pressures and expand the fluid through an expander to get some recoverable power which can be transmitted to the crankshaft. A schematic of the Rankine system can be seen in Fig. 1. The Rankine cycle efficiency is dependent on the temperature of the high temperature heat source (heat exchanger temperature in Fig. 1)

Corresponding author. E-mail address: [email protected] (V. Singh).

https://doi.org/10.1016/j.applthermaleng.2020.115157 Received 1 October 2019; Received in revised form 13 February 2020; Accepted 3 March 2020 Available online 04 March 2020 1359-4311/ © 2020 Elsevier Ltd. All rights reserved.

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Nomenclature BMEP CA50

Fuel MEP Fuel Mean effective Pressure HD Heavy Duty IMEPg Gross Indicated Mean Effective Pressure IMEPn Net Indicated Mean Effective Pressure LD Light Duty NOx Nitrogen Oxides PID Proportional Integral Derivitive Controller RTD Resistance Temperature Detector SI Spark Ignition WHR Waste Heat Recovery

Brake Mean effective Pressure Combustion timing in crank angle degree when 50% of the fuel is burned Crank angle degree Conventional Diesel Combustion Compression Ignition Chemiluminescense Detector Exhaust Gas Recirculation

CAD CDC CI CLD EGR

and the low temperature heat sink (condenser temperature in Fig. 1). Increased temperatures of the high temperature heat source translates to an increase in maximum achievable Rankine cycle efficiency for a given working fluid and a fixed condenser temperature. The exhaust gases, the exhaust gas recirculation (EGR) gases and the coolant are the three major sources of waste heat as per a study done on WHR viability for HD engines by Rijpkema et al. [8]. While the EGR gases and the exhaust flow have a higher temperature, their mass flow rates fluctuate throughout engine operation which could mean a reduced average Rankine cycle efficiency. The coolant on the other hand has more stable mass flow rates but has lower temperatures, reducing the WHR system efficiency. The aim of this study is to evaluate the use of high coolant temperatures in a Light Duty (LD) engine with respect to engine efficiency, exhaust gas enthalpy and the energy quality for the coolant. It also aims to maximize the recovery of waste heat from the coolant and study the response of the WHR system with changing operating conditions. Previous studies have investigated the effects of increased coolant temperatures. These works primarily look at increasing the combustion engine efficiency by reducing the temperature difference between the coolant channel walls and the coolant, thereby reducing the heat transfer. For an SI engine, the studies by Attar [9] and Mamun and Ehsan [10] show a decrease in indicated power with increasing coolant temperatures as the charge density and hence the volumetric efficiency within the combustion chamber is decreased. These studies however, go up to a maximum of 120 °C coolant temperature. They are also only applicable for SI engines, as for CI engines the fuel is not part of the charge inducted into the cylinder during the intake stroke, hence there is no loss in injected fuel in the cylinder with increased coolant temperature. For CI engines, studies were performed by Adler and Bandhauer [11] on a light duty diesel engine where the coolant temperature was swept from 90 °C to 150 °C. They saw a relative decrease in brake efficiency of up to 7.3% by increasing the coolant temperature. As per Adler and Bandhauer, this may be due to increased combustion durations at higher temperatures which also contribute to the higher exhaust temperatures. The total exergy as well as the coolant exergy showed a 20% to 40% increase over this interval. The exhaust

Exhaust Gas Out

Heat Exchanger

temperatures showed an increase of up to 100 °C for the highest load point run. However, the study was done for only low load conditions (4.3 bar BMEP). Here, also, the oil temperature was kept fixed at 80 °C, increasing the heat transfer to the oil. They also showed that the NOx emissions were largely unchanged when increasing coolant temperatures. Contrary to this, similar studies were done by Abdelghaffar [12] on increasing the coolant temperature for heavy duty engines where an increase in NOx emissions was observed. Another study on the effect was done by Slatar [13] for heavy duty (HD) engines. Here, the coolant temperature was varied up to 110 °C and 130 °C for the cylinder head and the liner respectively. It was seen that higher temperatures were preferred for the liner whereas lower temperatures were preferred for the cylinder head for increasing efficiency. However, Slatar states that the reliability of these results are low as the system noise was high due to unstable fuel injections. There is also work done on integrated cooling waste heat recovery circuits by Dingel et al. [14]. In this method, the coolant also acts as the working fluid for the Rankine cycle, being partly evaporated in the cooling channels of the engine. This has advantages over traditional cooling systems by maintaining more even metal temperatures in the engine and also improves the amount of waste heat that can be recovered from the coolant. In their studies, Dingel et al. found a reduction of the brake specific fuel consumption (BSFC) of up to 9.3%. However, this technology comes with the need to change the engine block design as the cooling channels face significantly higher pressures, being part of the Rankine cycle. Apart from the work done on elevated coolant temperatures, there are also studies that look at the recovery of the low temperature waste heat from an engine. A simulation based study was done by Singh et al. [15] for their work on a single cylinder Scania D13 heavy duty engine looking into recovering exhaust as well as coolant thermal energy. There, it was seen that the maximum recoverable power for a specific load-speed point is at a particular coolant temperature, due to the balance between the higher Rankine cycle recovery efficiency because of higher coolant temperatures and the reduced available heat due to reduced heat transfer to the coolant. However, the calculations for the WHR system were done using a Carnot efficiency to approximate the maximum recoverable power instead of detailed Rankine cycle simulations. Endo et al. [16] in their work also used higher coolant temperatures (189 °C) to study the WHR system for a light duty engine. The results showed a 13.2% relative increase in thermal efficiency of the engine. However, in this case, only a single coolant temperature was selected and the work does not characterise the change in recoverable power with coolant temperature. Another study was done by Fu et al. [17] where they evaluated four different working fluids to extract work from low temperature sources for a gasoline direct injection engine. Out of the four working fluids evaluated, R124 was taken as the best performing, showing the largest pressure ratio for the coolant temperatures used. The coolant temperature for the study was fixed at 91.9 °C. The brake efficiency was improved in the study by 12.1% when reaching the maximum allowable working fluid pressure of 16 bar.

Exhaust Gas In 3

2

Pump

Expander

1

Condenser

4

Fig. 1. Example of a Rankine Cycle System for Exhaust Waste Heat Recovery. 2

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Similarly, Dolz et al. [18,19] in their works also look at the waste heat recovery from a heavy duty diesel engine, evaluating 8 working fluids and multiple heat sources from the engine. They found the maximum power increment for the engine is obtained from a separated/binary cycle for the low temperature and high temperature heat sources. Using water and R245fa for the low temperature and high temperature cycles respectively, the authors found that a power increment of up to 19% could be achieved. Leduc et al. [20] in their work study low temperature WHR and its feasibility for internal combustion engines. The study is not as quantitative, instead highlighting some of the benefits of using such systems. Some of the advantages over high temperature WHR systems seen are lower pressure ratios, which means lighter system components and the relatively stable coolant temperatures over a driving cycle which further allows better control of the organic Rankine cycle. The system in development aims to improve the fuel economy by 2% to 3%. There are also multiple studies done on the low temperature WHR systems for marine engines and heavy duty diesel engines [21–25]. These studies look at different configurations for high temperature and low temperature WHR circuits and with different working fluids. For example, the work by Song et al. [21] for a marine diesel engine details the comparison of two different recovery circuits and 12 different working fluids. The best results were obtained for a separated recovery circuit system for high temperature sources and low temperature sources, showing up to 10.2% improvement in brake efficiency using R245fa and benzene as the working fluids for the cooling water and exhaust gases respectively. An additional source of low temperature WHR that has been studied is the engine lubricating oil itself. One study done by Soffiato et al. [26] studies the recovery of waste heat for a dual fuel marine engine application. Testing three different configurations for organic Rankine cycles (ORCs), they showed an improvement of 3.5% increase in power output producing up to 820.3 kW using the cooling water flows and the lubricating oil flow. The most optimum design for the WHR system was a two stage ORC circuit with R-236fa as the best performing working fluid out of the five refrigerants tested. Yang [27] in his work on marine diesel engines also looked at the lubricating oil as a source of waste heat. Here, the waste heat sources considered were the exhaust gas, the cooling water, scavenge air cooling water and the lubricating oil. The study analyses the performance of six working fluids for a transcritical Rankine cycle (TRC) recovery system. From a performance perspective, R152a provided the highest thermal efficiency and the highest power output out of all the fluids simulated.

Most studies which look at elevated coolant temperatures, use a maximum coolant temperature of 120 °C. These studies are also not concerned with looking at the recovery of waste heat but are more focused on reducing the heat losses from the combustion process and improving engine efficiency. For the research work described above looking at low temperature WHR, there is very little work done on optimising the engine along with the WHR system. Generally, the engine parameters such as the coolant temperature are treated as constant without trying to maximise the exergy of the coolant flow. These studies also generally compare 6 or fewer working fluids per heat source and find an optimum working fluid without comparing a larger group of test fluids in their simulations. The studies also tend to optimise the WHR loop for a specific engine operation point and not for changing engine loads and speeds. To address these knowledge gaps, this paper aims to evaluate both the engine and the WHR system at different engine operating conditions to study the response of the entire system with changing engine load and speed. The work described also aims to optimise the coolant temperature to maximise recoverable power while also evaluating 48 different working fluids to get a better picture of which working fluids perform better for both low temperature and high temperature heat sources. 2. Methodology To study the effect of elevated coolant temperatures on the complete system (engine and waste heat recovery) a combination of experimental work and simulations was used. Coolant temperature sweeps were first done experimentally on a Volvo D4 LD engine. The results from the temperature sweeps were then used for detailed Rankine cycle simulations using Modelica. Here, 48 different working fluids were simulated in a dual Rankine cycle setup to see the recoverable power. The following sections explain the experimental setup and the simulation setup for the study. 2.1. Experimental setup The engine used for the study was a 4 cylinder Volvo VED4 light duty engine. The engine does not have a turbocharger, being replaced by a compressed air supply and a valve to give backpressure on the exhaust side. This was done so as to have independent control of the intake pressure and backpressure for the engine. The inlet air temperature to the engine is controlled using an air heater. The engine schematic is shown in Fig. 2.

Fig. 2. Volvo D4 Test Cell Layout. 3

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The oil cooling circuit for the engine was modified so as to cool it using tap water instead of using the coolant. This was done to maintain a lower engine oil temperature and prevent the engine from seizing due to possibly excessively low oil viscosities at higher coolant temperatures. The oil used for the experiments was SAE 10 W-60. A Macnaught MX12 Oval gear flow meter was used to measure the flow rate of the oil with RS Pro Type K thermocouples to measure the inlet and outlet temperatures of the oil from the heat exchanger. For emissions, an AVL AMA i60 emissions system is used to measure CO, CO2, NOx, NO and THC from the exhaust gases. The CO and CO2 are measured using an Infrared Detector with the Total Hydrocarbons being measured using a Flame Ionisation Detector. The O2 concentration is determined by a Paramagnetic Detector and the NO and NOx concentrations are measured by a Chemiluminescense Detector. It should be noted here that for such a system (shown in Fig. 3) to be in actual application on a passenger vehicle, the vehicle would require some modifications. The addition of an oil cooler, valves and PID controllers for the accurate control of coolant and oil temperatures would be needed. Resizing of the radiator could also be done to due to the reduced cooling requirements for the coolant.

Table 1 Volvo D4 engine specifications. Engine Parameter

Value

Displacement (L) Compression Ratio Bore × Stroke (mm) Connecting Rod Length (mm) Number of Valves/Cylinder Inlet Temperature Fuel Combustion Mode

2.0 13.7:1 82 × 93.2 147 4 50 °C Diesel MK1 CDC

The specifications for the engine are detailed in Table 1. The oil and coolant cooling circuits for the test setup are shown in Fig. 3. The coolant in the engine is pumped by a belt driven water pump. Hence the coolant flow rate is engine speed dependent. The coolant is circulated through to a plate heat exchanger which is further cooled by circulating tap water. The tap water flow is regulated using a PID controlled valve to modify the cooling capacity of the heat exchanger. The circuit also has a bypass valve installed to have some of the coolant flow bypass the plate heat exchanger. This is done as at higher coolant or oil temperatures the tap water in the heat exchanger starts boiling which increases the heat transfer and restricts the coolant and oil temperature at a certain value. The bypass valve allows some of the coolant or oil to bypass the heat exchanger making sure that the tap water does not start boiling. At increased coolant or oil temperatures, the bypass valve is opened more. The coolant used for the experiments was pure ethylene glycol. Two TE Connectivity Resistance Temperature Detectors (RTDs) were used to measure the temperature in and out of the heat exchanger (after the bypass, allowing sufficient mixing time) with a Sandhurst Instruments LX-25 turbine flow meter to measure the coolant flow rate. For the calculation of the efficiencies, the engine torque for brake efficiency, the in-cylinder pressure for indicated efficiency, and the mass flow rate of fuel is measured. The engine torque is measured using an HBM T40B torque transducer mounted on the propeller shaft from the dynamometer. The in-cylinder pressure measurement is done using an AVL GH 14D pressure transducer. The cylinder pressure is pegged using the intake pressure measurement. For the fuel flow measurements, the supply fuel tank for the engine is mounted on a Sartorius weighing scale to measure the rate of change of fuel mass, obtained by linear fitting of the fuel weight measured. The sensor accuracy and the uncertainty in measured values is discussed in the following sections.

2.2. Operating conditions The operating conditions run for the engine are detailed in Table 2. Two loads and two speeds were tested with the engine. The temperature sweep was attempted from 80 °C to 160 °C for all operating points. However, as the coolant temperature is increased, the heat rejected to it is reduced until it becomes negligible i.e. the cylinder heat is not transferred to the coolant any more. At the lower load point tested (10 bar IMEPg), the heat rejected to the coolant became negligible at a coolant temperature of 120 °C. Hence at lower loads, as there is not much heat rejected and as the metal temperatures in the engine are lower, the maximum coolant temperature reached is also lower. For the coolant temperature sweep, only the coolant outlet temperature from the engine block was controlled, the coolant inlet temperature to the block was allowed to vary depending on the changes in heat transfer. Two different loads were selected to understand the variation of recoverable power with load and how the optimum coolant temperature is changed. The two different speeds were taken to understand the system response with changing mass flow rates of air, fuel, coolant and oil and the resulting changes in heat transfer rates to the coolant. No external EGR was used in the experiments, however the backpressure was kept the same as the intake pressure for the engine to

Fig. 3. Cooling circuits for Volvo D4 engine. 4

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air flow with the increase of coolant temperature due to increased in-cylinder gas temperatures. This is slightly mitigated by maintaining a constant air temperature at the intake port. However, a reduction of air mass flow rates of up to 3.1% was seen when increasing the coolant temperature from 80 °C to 160 °C. The combined effect of the reduced air and fuel flow changes the lambda value by less than 1%. Hence, the change in lambda due to increasing coolant temperatures was not considered to affect the exhaust gas temperature significantly. For the data itself, each point has 150 cycles of combustion data saved for analysis.

Table 2 Operating points used for experimental campaign. Point 1 Load (Bar IMEPg) Engine Speed (RPM) Indicated Power (kW) Injection Pressure (Bar) CA50 (ATDC) Fuel Temperature Sweep (°C) Lambda Intake Temperature (°C) Back Pressure Oil Temperature (°C) Oil Type External EGR

Point 2

10 1200 1800 20 30 900 10 Diesel MK1 80–120 1.9 50 ~= Intake Pressure 110 SAE 10 W-60 0%

Point 3

Point 4

15 1200 30

1800 45

15 80–160 1.6

2.3. Experimental uncertainty and error analysis For the experiments, the variation in the controlled parameters is shown in Table 3. This is shown as the maximum variation in the controlled parameter from the target value for all of the coolant temperature sweeps. For the sensors used, the details and the uncertainty for the whole measurement range is detailed in Table 4. For the measured results, the error analysis for the calculated parameters: heat transfer to the oil, heat transfer to the coolant and exhaust enthalpy is shown in Fig. 4 over the coolant temperature sweep range. The errors in brake and friction are not shown as they are negligible compared to the other errors shown in the figure. The error in exhaust gas enthalpy calculations is low due to the high exhaust gas temperatures and the relative accuracy of the thermocouples in determining the exhaust gas temperatures and hence, enthalpy. The absolute values for the heat loss to each media (exhaust gas, coolant or oil) is calculated through Equation (1).

Table 3 Variations in controlled parameters for experimental study. Parameter

Value(s)

Max. Variation from Target

Engine Speed Coolant Outlet Temperature

1200 RPM, 1800 RPM 80 °C, 120 °C, 140 °C, 160 °C 110 °C, 80 °C 50 °C 900 bar 10 °ATDC, 15 °ATDC

± 2 RPM ± 2 °C

Oil Temperature Inlet Temperature Rail Pressure CA50

± 1 °C ± 1 °C ± 10 bar ± 0.1 °ATDC

simulate a turbocharger and hence some internal EGR can be expected. The intake temperature is regulated to 50 °C using a PID controlled air heater to maintain a constant intake port temperature to the cylinders. The injection pressure was maintained at 900 bar to keep the pressure rise rate for all the load-speed points at lower than 10 bar/CAD. The oil sump temperature was maintained at 110 °C using the external oil cooling circuit. For a temperature sweep at a specific load-speed point, the fuel injection duration was kept constant and the aim was to vary the injection timing to maintain a constant CA50 for the sweep. However, it was seen that the CA50 was largely unaffected by the coolant temperature, which is to be expected in mixing controlled diffusion combustion, hence the start of injection was also constant for the temperature sweeps. The back pressure for the experiments was kept equal to the intake pressure. Here, the additional backpressure from the heat exchanger for the WHR system is not accounted for as that is highly dependent on the design of the heat exchanger itself which is not implemented in the experimental setup. As the injection timing and duration was kept constant, consequently the volumetric flow rate of the fuel is also constant, however the fuel mass flow rate is reduced due to higher metal temperatures in the engine. This led to a reduction of mass flow rate of the fuel due to thermal expansion of up to 2.6% going from a coolant temperature of 80 °C to 160 °C. Hence the Fuel MEP is reduced at higher coolant temperatures. There is also a reduction in

(1)

Heat Loss (HL) = mCp T

Here, the m , Cp and T denote the mass flow rate, specific heat and the temperature change of each heat loss medium. For the exhaust gas, the T denotes the temperature change from the exhaust gas temperature at the exhaust port to the inlet temperature (50 °C). The error in this computation is calculated using Equation (2).

HL = |HL|·

m m

2

+

T T

2

(2)

In the equation above, the term HL is used to describe the heat loss. HL , m and T is the error in the calculation or measurement of the respective values of heat loss, mass flow rate and temperature difference for any media (exhaust gas, coolant or oil). Here, Cp is assumed as T is further calculated using Equation (3). a constant. The term

T=

(3)

T12 + T22

T1 and T2 are the errors in the measurement of the two temperatures T1 and T2 in determining T . As can be seen in Fig. 4, the errors in computation of the heat transfer to the coolant is decreased at higher coolant temperatures due to the reduction in heat transfer to the coolant The errors in computation of heat transfer to the oil is increased at higher coolant temperatures due to the increased heat transfer to the oil. The error margins

Table 4 Sensor types and accuracies. Sensor

Type

Accuracy

Coolant Inlet and Outlet Temperature

Class AA RTD

Oil Inlet and Outlet Temperature Oil Flow Rate Sensor Coolant Flow Rate Sensor Cylinder Pressure Sensor Exhaust Temperature Sensor Torque Sensor

Thermocouple Oval Gear Flow Meter Turbine Flow Meter Piezoelectric Pressure Sensor Thermocouple Strain Gage Sensor

± 0.24 °C at 80 °C reading ± 0.37 °C at 160 °C reading ± 1.5 °C ± 0.5% ± 0.5% ± 1.25 bar ± 1.5 °C ± 0.03%

5

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Fig. 4. Experimental error for computed values.

are high with respect to the absolute values, as the flow rates for the coolant and the oil to the engine were maintained, causing the maximum temperature differential between the coolant/oil inlet and outlet of the engine to be below 10 °C. This is to prevent uneven metal temperature distribution which could negatively affect the engine. However, with these values, trends can still be observed in the heat transfer to the two fluids – the coolant and oil. For example, the heat transfer to the coolant is reduced from 19 kW to 1.3 kW when increasing the coolant temperature from 80 °C to 160 °C (shown in Results section). This trend in heat transfer can be compared to the errors in computation at 5.2 and 1.3 kW at 80 °C and 160 °C respectively.

thermodynamic cycle models used were developed in the Modelica programming language with CoolProp being used for the working fluid properties and the ExternalMedia interface being used to connect the two. The reference conditions, boundary conditions and constraints for the simulation are detailed in Table 5. The pump and expander efficiency are set to 60% based on works by Rijpkema et al. [28] and previously reported experimental values. The maximum pressure was set to 125 bar and added constraints were placed to avoid operation near the thermodynamic critical pressure and temperature (pcrit, Tcrit). The pumps and expanders are modeled using a simple isentropic efficiency for expansion or compression and the heat exchangers are modeled using a pinch point analysis. For the simulations, water is removed from the composition of the exhaust gases as the condensation of water can lead to erroneous results, showing a much higher available heat from the exhaust gases than in practical application. Additionally, CoolProp does not have ethylene glycol

2.4. Simulation setup The dual loop Rankine cycle system (shown in Fig. 5) was simulated with the two recovery loops being for the exhaust gases and the coolant. The Pump

Coolant In

Table 5 Reference conditions, boundary conditions and constraints for Rankine cycle simulations.

Evaporator

Condenser

Intake

Reference and Boundary Conditions

Engine

Exhaust Coolant Out

Ambient Temperature (°C) Ambient Pressure (bar) Pump/Expander Efficiency (%) Pump Vapor Quality In

Evaporator

25 1.013 60 0

Constraints

Expander

Max. Evaporation Pressure (bar)

Pump

Expander

Min. Condensation Temperature (°C) Min. Condensation Pressure (bar) Pinch Point (°C) Max. Cycle Pressure (bar) Max. Superheating (°C)

Condenser

Fig. 5. Dual loop Rankine cycle simulation setup. The data for the engine, coolant flows and intake and exhaust flows are taken from experiments. 6

min

{

125 0.9·pcrit

50 1.013 5 125 100

}

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IMEPg, 1200 RPM and 110 °C oil temperature. Here it can be seen that the heat transfer to the oil is increased at higher coolant temperatures, with the combined heat transfer to the oil and coolant being reduced. The ‘Remaining’ term in the energy balance is the energy remaining from the fuel energy injected per cycle that is not accounted for in the other terms. This includes all other kinds of losses, but mainly the heat loss from the engine to the ambient air in the test cell. It can be seen that the remaining power is also increased at higher coolant temperatures, due to higher engine metal temperatures which increases the convective heat loss to the ambient air. It should be noted here that it is assumed that the power taken from the friction is given to the oil and acts to heat up the oil. Hence the frictional power is subtracted from the heat transfer to the oil when doing the energy balance analysis to show only the direct heat loss from the combustion chamber to the oil. The exhaust, brake and friction energy look largely unaffected from the energy balance diagrams, however they show significant trends on closer examination, as shown later. These trends in the energy balance with increasing coolant temperature is seen for all the operating points tested (for 110 °C oil temperature). The net power loss due to heat transfer, taking into account the heat losses to the ambient as well as the losses to the oil and coolant is shown in Fig. 7. Here, it can be seen that the overall heat transfer decreases by up to 4 percentage points of the fuel energy for each case. The reduction in net heat transfer is not as much at higher engine speeds. This is because of the lower heat transfer to the coolant per cycle at higher speeds due to the reduced time per cycle. Hence, the reduction in heat transfer is lower with increasing coolant temperature at higher engine speeds.

Fig. 6. Energy balance for temperature sweep at 15 bar load, 1200 RPM, 110 °C oil temperature.

in the list of fluids. To compensate for this, water is used as the coolant for the Rankine cycle simulations with the mass flow rate modified to scale with the ratio of specific heat of water to ethylene glycol. Thus, the heat flow through the heat exchanger is kept the same as for ethylene glycol. 3. Results

3.1.2. Engine efficiency The brake efficiency for the four operating points can be seen in Fig. 8. Here, the brake efficiency is seen to increase as the coolant temperature increases. This is due to an increase in indicated efficiency due to lower heat transfer from the combustion chamber (shown in

3.1. Experimental results 3.1.1. Energy balance Fig. 6 shows the energy balance for the temperature sweep at 15 bar

Fig. 7. Net heat loss for 110 °C oil temperature points shown as a fraction of the fuel energy. 7

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Fig. 8. Engine brake efficiency with increasing coolant temperatures.

Fig. 9) as well as increased mechanical efficiency with increased coolant temperature due to lower oil viscosities (shown in Fig. 10). The mechanical efficiency is calculated using the formula in Equation (4). Here, the BMEP is the brake mean effective pressure and IMEPn is the net indicated mean effective pressure. mech

=

BMEP IMEPn

At higher engine speeds the gain in mechanical efficiency is higher for increasing coolant temperatures. This is because of the higher frictional losses at higher engine speeds, making it more sensitive to changes in oil viscosity. At the lower engine speed of 1200 RPM, the change in mechanical efficiency is low or insignificant. 3.1.3. Exhaust temperature and enthalpy With increasing coolant temperatures, higher exhaust gas temperatures are also seen, as shown in Fig. 11. This is to be expected due to the lower heat transfer to the coolant. At higher loads, the exhaust

(4)

At higher engine loads the brake efficiency does not increase any further beyond a coolant temperature of 120 °C.

Fig. 9. Gross Indicated efficiency with increasing coolant temperatures. 8

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Fig. 10. Mechanical efficiency for increasing coolant temperatures.

gas temperature is increased over 20 °C. This rise in temperature could be increased further by using higher oil temperatures and reducing the heat transfer losses to the ambient air using insulation for the engine. Fig. 12 shows the changes in exhaust gas enthalpy for the engine. Here, it can be seen that while the exhaust gas enthalpy increases

generally with increasing coolant temperature, the changes are not as significant as the changes in exhaust temperature. This is due to the reduced mass flow rate of air at higher temperatures as the volumetric efficiency of the engine is reduced. This is despite maintaining constant inlet pressure and inlet air temperature at the ports. The expansion of

Fig. 11. Exhaust gas temperatures (measured at ports) for changing engine temperatures. 9

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Fig. 12. Change in exhaust enthalpy for increasing coolant temperatures.

gases during the intake stroke due to the hotter in-cylinder temperatures shows a steady decrease in air flow rate and hence the exhaust enthalpy is a balance between the rising exhaust temperature and the lower mass flow rate.

3.1.4. NOx emissions Fig. 13 shows the change in NOx emissions with elevated temperatures. This is as expected, as the higher coolant temperatures also mean higher in-cylinder temperatures and hence overall higher NOx emissions. At the 10 bar IMEP, 1800 RPM point the NOx emissions are

Fig. 13. NOx emissions for increasing coolant temperatures.

10

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Fig. 14. Maximum power output from the (top) exhaust and (bottom) coolant recovery loop for the different working fluids tested at 15 bar gross IMEP, 1800 RPM operating point.

reduced at higher temperatures. This is due to the reduced air flow and hence exhaust gas flow at higher temperatures, thus reducing the emission levels. If seen in parts per million (independent of exhaust mass flow rate), the NOx levels are always increasing with increasing coolant temperature. It was also seen that the NOx emissions were higher at lower engine loads. This is primarly due to the earlier combustion phasing which increases the cylinder temperature as well as the lower power produced. Lower NOx emissions are also seen at higher engine speeds due to the reduced residence time of the combustion gases in the cylinder. At lowest load and speed, the NOx emissions are increased by up to 0.9 g/kWh.

3.2. Simulation results The simulations for the Rankine cycle recovery system were done in Modelica using 48 working fluids as the possible candidates for the recovery cycles. The data for the exhaust gas temperature, pressure, mass flow rate and composition were taken from the experiments as inputs for the high temperature WHR simulations. Similarly, the coolant temperatures (inlet and outlet) and mass flow rate were used for the low temperature WHR simulations. A scatter plot showing the maximum power output against the pressure ratio across the expander for all the working fluids tested with the coolant and exhaust for the 15 bar, 1800 RPM (high load, high speed) operating point can be seen in Fig. 14. 11

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Fig. 15. Recoverable Power from the expander for the best performing working fluids for each operating point. The blue line shows the power recovered from the coolant and the orange line represents power recovered from the exhaust.

For the coolant, cyclopentane provided the highest power output out of all the fluids tested with up to 1 kW delivered from the expander. Cyclopentane also provided the highest power output for the other operating points tested. In the other cases the power output from the Rankine cycle is reduced as the heat rejected is reduced at lower engine loads and speeds, however the best performing fluids are similar across the operating points. This is because the performance is highly dependent on the temperature of the heat source and the working fluid which matches with those temperatures. For the exhaust, the maximum power output delivered from the expander is up to 4.25 kW for methanol as the working fluid. Here, the temperaure of the exhaust gases is changed with the load and speed, hence the best performing fluid also changes. At 10 bar gross IMEP load, the best working fluid was found to be acetone and for the 15 bar points, it was found to be methanol. However, even at the lower load point, methanol produces only 0.05 kW lower power and performs increasingly better the higher in load and speed the engine is operated. In the case of the exhaust WHR system, the pressure ratio reached is up to 60:1. In practical application, this could be done using a set of expanders (piston or scroll) in series. This would change the expander design and expander efficiencies based on the operating conditions of each individual expander. However, the expander modeling itself is considered outside the scope of the study and a fixed expander efficiency was assumed for the simulations. For the coolant and exhaust the power output with increasing coolant temperature is shown in Fig. 15 for the best performing working fluids for each operating point. It can be seen that with higher engine loads and speeds the optimum coolant temperature is higher for maximising recoverable power from the coolant. At 10 bar load, the highest recoverable power Is seen at 100 °C coolant temperature, whereas at 15 bar load the highest coolant recoverable power is at 120 °C. This is due to more available heat from

the coolant at higher temperatures. The optimum coolant temperature for maximizing recoverable power is a balance between the higher recovery efficiency of the Rankine cycle at higher coolant temperatures and the lower available energy due to lower heat rejection to the coolant. The recoverable power from the exhaust increases with increasing coolant temperatures, however it plateaus at higher engine speeds and higher coolant temperatures as the reduction in air density reduces the exhaust enthalpy and limits the amount of recovered power. The increase in exhaust recoverable power is however, not as significant as the changes in coolant recoverable power. This is because the changes in exhaust gas temperatures and exhaust gas enthalpy is lower. It is also seen that the maximum coolant recoverable power is higher at lower engine speeds. This is due to the higher heat available in the coolant at lower engine speeds. Fig. 16 shows the combined change in system efficiency with increasing coolant temperatures for all the operating points. The system efficiency with WHR is calculated as shown in Equation (5).

=

Brake Power + Recovered Power ·100 Fuel Power

(5)

The figure shows a plot of the efficiency when using only the recoverable power from the coolant in Equation (4) and when using the recovered power from both the coolant and the exhaust gases to show the effects of each heat source. Here, it can be seen that the implementation of the WHR system shows a significant improvement over the base engine efficiency. The maximum improvement in system efficiency is up to 5.2 percentage points for the higher load cases (which can be further seen in Fig. 17). Of this, up to 1.7 percentage points is from the power recovered from the coolant.

12

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Fig. 16. Overall system efficiency with and without the use of a waste heat recovery system.

Fig. 17. Efficiency improvements in percentage points from the implementation of a waste heat recovery system. 13

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4. Conclusions A Volvo VED4 engine was run with varying coolant temperatures to study the effect on recoverable power and the engine characteristics. From the experimental studies it was observed that:

•

• An increase in coolant temperature from 80 °C to 160 °C for the

• •

engine resulted in an increase in brake efficiency of the engine by up to 1 percentage point. This was the result of an increase in gross indicated efficiency due to reduced heat transfer losses and an increase in mechanical efficiency due to reduced oil viscosity of the oil on the cylinder liner. With increased coolant temperatures, the exhaust gas enthalpy and temperatures are also increased. However, there is a reduction in mass flow rate of air in the combustion chamber at higher coolant temperatures due to reduced density. Increased coolant temperatures from 80 °C to 160 °C were seen to increase the NOx emissions by up to 0.9 g/kWh due to higher incylinder gas temperatures.

•

Declaration of Competing Interest The authors declare that they have no known competing financial interests or personal relationships that could have appeared to influence the work reported in this paper.

From the Modelica simulations for a dual loop Rankine cyle WHR system the following conclusions can be drawn:

• To

•

•

• •

maximise recoverable power from the coolant, an optimum coolant temperature is observed which is dependent on the engine operating conditions. There is an optimum temperature for each operating point because at higher coolant temperatures the efficiency of the low temperature WHR system is increased however there is lower rejected heat to the coolant which can be used by the Rankine cycle. So there is a coolant temperature at which the sytem as a whole has maximum brake efficiency. The optimum coolant temperature is dependent on engine load and speed. And engine operating point which has higher heat losses to the coolant would shift the optimum coolant temperature higher. This is because the coolant temperature at which the heat rejected to coolant is low enough to reduce the Rankine cycle output power is at a higher temperature. Out of the 48 working fluids simulated for the low temperature WHR system, cyclopentane showed the best performance in terms of recoverable power for all engine loads and speeds tested. An efficiency improvement of up to 1.7 percentage point in system brake efficiency was seen from only the coolant WHR with the use of cyclopentane and at the optimum coolant temperatures. For the exhaust, from the 48 different working fluids tested, the best performing working fluids were methanol or acetone depending on engine operating conditions. At engine operating points with higher heat rejection rate to the coolant, methanol was seen to perform better and recover more power from the exhaust waste heat. The combined dual loop WHR system shows an improvement of up to 5.2 percentage points from the base engine using both an optimised coolant temperature and selecting the working fluid appropriate to the heat source.

Acknowledgements The authors would like to acknowledge the Swedish Energy Agency, Scania CV, Volvo Personvagnar AB, AB Volvo, Titan X, Gnutti Carlo Sweden AB and IAV for their funding and support in the waste heat recovery project (Project Number: 32599-3). References [1] H. Teng, G. Regner, C. Cowland, Waste Heat Recovery of Heavy-Duty Diesel Engines by Organic Rankine Cycle Part I: Hybrid Energy System of Diesel and Rankine Engine, SAE Technical Paper 2007-01-0537, 2007. [2] N. Abani, N. Nagar, R. Zermeno, M. Chiang, et al., Developing a 55% BTE Commercial Heavy-Duty Opposed-Piston Engine without a Waste Heat Recovery System, SAE Technical Paper 2017-01-0638, 2017. [3] J. O'Connor, M. Borz, D. Ruth, J. Han, et al., Optimization of an advanced combustion strategy towards 55% BTE for the Volvo SuperTruck Program, SAE Int. J. Engines 10 (3) (2017). [4] V. Manente, B. Johansson, P. Tunestal, Partially Premixed Combustion at High Load using Gasoline and Ethanol, a Comparison with Diesel, SAE Technical Paper 200901-0944, 2009. [5] A. Legros, L. Guillaume, M. Diny, H. Zaidi, V. Lemort, Comparison and impact of waste heat recovery technologies on passenger car fuel consumption in a normalized driving cycle, Energies 7 (8) (2014) 5273–5290. [6] D.D. Battista, R. Cipollone, Improving engine oil warm up through waste heat recovery, Energies 11 (2018). [7] R. Cipollone, D.D. Battista, M. Mauriello, Effects of oil warm up acceleration on the fuel consumption of reciprocating internal combustion engines, Energy Procedia 82 (2015) 1–8. [8] J. Rijpkema, K. Munch, S. Andersson, Thermodynamic potential of Rankine and flash cycles for waste heat recovery in a heavy duty Diesel engine, Energy Procedia 129 (2017) 746–753. [9] A.M. Attar, The Effects of Coolant Temperature on Spark Ignition Engine Performance, Western Michigan University, Kalamazoo, 2009. [10] M. Mamun, M. Ehsan, Effect of coolant temperature on performance of a SI engine, in: 4th International Conference on Mechanical Engineering, Dhaka, 2001. [11] J. Adler, T. Bandhauer, Performance of a diesel engine at high coolant temperatures, J. Energy Res. Technol. 139 (2017). [12] W.A. Abdelghaffar, M.M. Osman, M.N. Saeed, A.I. Abdelfatteh, Effects of coolant temperature on the performance and emissions of a diesel engine, in: ASME 2002 Internal Combustion Engine Division Spring Technical Conference, Rockford, 2002. [13] A. Slatar, Influence of Coolant Temperature and Flow on Engine Efficiency, Department of Aeronautical and Vehicle Engineering, KTH Royal Institute of Technology, Stockholm, 2015. [14] O. Dingel, D. Luederitz, T. Arnold, Investigation of an ORC System with Integrated Phase Change Engine Cooling, in: Proceedings of the 5th International Seminar on ORC Power Systems, Athens, 2019. [15] V. Singh, P. Tunestal, M. Tuner, A Study on the Effect of Elevated Coolant Temperatures on HD Engines, SAE Technical Paper 2017-01-2223, 2017. [16] T. Endo, S. Kawajiri, Y. Kojima, K. Takahashi, et al., Study on Maximizing Exergy in Automotive Engines, SAE Technical Paper 2007-01-0257, 2007.

From the conclusions drawn above in the study, the following can be said regarding the use of low temperature WHR in passenger vehicles:

• For the use of coolant as a heat source for Rankine cycle WHR higher

•

performance for the mass flows and temperature range for the light duty diesel engine studied and could be considered as a potential working fluid. In addition to methanol and acetone showing better performance than the other working fluids for the recovery of exhaust gas waste heat, they also perform well for the recovery of waste heat from the coolant. Hence, they could be considered good candidates for the implementation of a Rankine cycle WHR system using a single working fluid and and both heat sources (low temperature and high temperature) in a single loop. The increase in NOx emissions from the engine needs to be taken into account when using higher coolant temperatures. This can be mitigated through the use of Selective Catalytic Reduction (SCR) systems, however this translates to additional costs in production. Another method is through the use of EGR, however this means changing the combustion characteristics within the combustion chamber.

temperatures show a significant increase in recoverable power from the engine. An optimum temperature for each load speed point can be determined and mapped for the engine operating range. Furthermore, a coolant temperature controller can be implemented to keep the coolant temperature at the optimum point. In most studies of low temperature WHR, the main candidates for the working fluid for coolant WHR seen in literature are generally refrigerants. However cyclopentane showed consistently good 14

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