Characterization of Engine Performance with Biodiesel
Fuels
by
Timothy Philip Guider
A Thesis
Presented to the Graduate and Research Committee
of Lehigh University
in Candidacy for the Degree of
Master of Science
in
Mechanical Engineering
Lehigh University
(December 2008)
This thesis is accepted and approved in partial fulfillment of the requirements
for the Master of Science
Date
Dr. Sudhakar Neti, Thesis Advisor
Dr. Gary Harlow, Chairperson of Department
ii
Acknowledgments
First, I would like to thank Dr. Sudhakar Neti for the opportunity to conduct
research in this growing and evolving field as well as for his guidance throughout the
last year and a half.
I‟d also like to thank Fred Betz for his help in using his CCHP system at
Carnegie-Mellon University (CMU). Dr. David Archer and Dr. John Wiss were also
extremely helpful during the weekend trips to CMU.
This research could not have been done if it weren‟t for Dick Towne, head of
the machine shop in Packard Lab. Dick helped setup the instrumentation for the
single-cylinder and gave me permission to run “unknown” fuels in his diesel engine.
I appreciate Justin Christenson‟s help in the engines lab with writing labView
VI‟s and helping me install/remove any parts of the engine at the drop of a hat.
Also, I‟d like to thank Jason Slipp for creating the biodiesel from the waste
vegetable oil and letting us use a good amount of it.
Thanks to Tim Nixon for his help with the DAQ and electrical work.
Thanks to YiJun Yang, my office mate, for putting up with me constantly
asking for his input on all subjects pertaining to this thesis.
And thanks to Brian Holder and Erony Whyte for their time put into
reading/reviewing my thesis.
I‟d also like to thank my family for standing by me and encouraging me
throughout graduate school. And of course, the biggest thanks to my girlfriend Jill for
her patience and support throughout the last year and a half.
iii
Table of Contents
Acknowledgments........................................................................................................ iii
Table of Contents ......................................................................................................... iv
List of Tables .............................................................................................................. vii
List of Figures ............................................................................................................ viii
ABSTRACT .................................................................................................................. 1
CHAPTER 1 ................................................................................................................. 2
INTRODUCTION ........................................................................................................ 2
1.1
Problem Statement ........................................................................................ 4
1.2
Scope of Present Work.................................................................................. 5
1.3
Outline of Thesis ........................................................................................... 5
CHAPTER 2 ................................................................................................................. 6
BACKGROUND .......................................................................................................... 6
CHAPTER 3 ............................................................................................................... 13
BIODIESELS .............................................................................................................. 13
3.1
Overview ..................................................................................................... 13
3.1.1
Organic Chemistry .............................................................................. 13
3.1.2
Feedstock ............................................................................................ 15
3.1.3
Alcohol ................................................................................................ 15
3.1.4
Catalyst ............................................................................................... 16
3.1.5
Neutralizer........................................................................................... 17
3.1.6
Glycerol............................................................................................... 17
iv
3.2
Waste Vegetable Oil (WVO) ...................................................................... 18
3.3
Soy Biodiesel .............................................................................................. 23
CHAPTER 4 ............................................................................................................... 24
SINGLE-CYLINDER DIESEL ENGINE .................................................................. 24
4.1
Experimental Setup ..................................................................................... 24
4.1.1
Engine ................................................................................................. 24
4.1.2
Load Cells ........................................................................................... 27
4.1.3
Dynamometer ...................................................................................... 27
4.1.4
Resolver .............................................................................................. 29
4.1.5
Manometer and Pulse-damping Drum ................................................ 30
4.1.6
Pressure Transducer ............................................................................ 32
4.1.7
Thermocouples .................................................................................... 33
4.2
Procedure and DAQ .................................................................................... 33
4.3
Results ......................................................................................................... 36
4.3.1
P-ν diagrams ....................................................................................... 36
4.3.2
Brake Mean Effective Pressure (BMEP) ............................................ 43
4.3.3
Brake Horsepower (BHP) ................................................................... 43
4.3.4
Mechanical Efficiency ........................................................................ 45
4.3.5
Brake Specific Fuel Consumption (BSFC) ......................................... 47
4.3.6
Brake Thermal Efficiency ................................................................... 50
4.3.7
Volumetric Efficiency......................................................................... 52
4.3.8
Net Exhaust Temperature ................................................................... 56
4.3.9
Potential CHP Efficiency .................................................................... 58
v
CHAPTER 5 ............................................................................................................... 62
CCHP TESTS WITH A FOUR-CYLINDER DIESEL ENGINE .............................. 62
5.1
Experimental Setup ..................................................................................... 64
5.1.1
Engine ................................................................................................. 65
5.1.2
Turbocharger ....................................................................................... 67
5.1.3
Steam System ...................................................................................... 76
5.1.4
Domestic Hot Water ........................................................................... 79
5.2
Procedure and DAQ .................................................................................... 81
5.3
Experimental Results .................................................................................. 84
5.3.1
Engine ................................................................................................. 84
5.3.2
Turbocharger ....................................................................................... 86
5.3.3
Steam System ...................................................................................... 94
5.3.4
Domestic Hot Water ........................................................................... 96
CHAPTER 6 ............................................................................................................... 99
RESULTS AND DISCUSSION ................................................................................. 99
REFERENCES ......................................................................................................... 107
APPENDIX ............................................................................................................... 110
VITA ......................................................................................................................... 113
vi
List of Tables
Table 1: Fuel Property Comparison of Vegetable Oils with No. 2 Diesel Fuel ........... 3
Table 2: Some Fuel Properties of Various Biodiesels .................................................. 4
Table 3: Soy Biodiesel Properties ............................................................................... 23
Table 4: Yanmar Engine Specifications ..................................................................... 24
Table 5: Calculated Values from PV-diagrams .......................................................... 38
Table 6: Specifications for John Deere 4024TF270 ................................................... 66
Table 7: Vaporphase Steam Generator Specifications................................................ 77
Table 8: Heat Exchanger Specifications ..................................................................... 80
Table 9: Measured Engine Data vs. Manufacturer Specifications .............................. 85
Table 10: Average Turbocharger Results for No. 2 Diesel ........................................ 89
Table 11: Average Turbocharger Results for Soy Biodiesel ...................................... 89
Table 12: CCHP Results with No. 2 Diesel Fuel ...................................................... 103
Table 13: CCHP Results with Soy Biodiesel ............................................................ 103
Table 14: Turbocharger Results for No. 2 Diesel ..................................................... 105
Table 15: Turbocharger Results for Soy Biodiesel ................................................... 105
vii
List of Figures
Figure 1: Fuel Consumption Rate and BSFC vs. RPM, Lin (2006) ............................. 7
Figure 2: Brake Thermal Efficiency and Exhaust Gas Temperature vs. RPM, Lin
(2006) ............................................................................................................................ 8
Figure 3: CO, CO2, and NOx Emission Indices vs. RPM, Lin (2006) .......................... 9
Figure 4: Thermal Efficiency vs. BMEP, Agarwal (2007) ......................................... 10
Figure 5: BFSC vs. BMEP, Agarwal (2007) .............................................................. 11
Figure 6: Exhaust Temperature vs. BMEP, Agarwal (2007) ...................................... 11
Figure 7: Transesterification Reaction ........................................................................ 14
Figure 8: Titration ....................................................................................................... 19
Figure 9: Biodiesel Mixture throughout Entire Process ............................................. 20
Figure 10: Recycled Grease Manufacturing System .................................................. 21
Figure 11: Schematic of Recycled Grease Manufacturing System ............................ 22
Figure 12: Single-Cylinder Test Setup ....................................................................... 26
Figure 13: Load Cell ................................................................................................... 27
Figure 14: Dynamometer Load Control System ......................................................... 28
Figure 15: Setup for Magnetic Pickup ........................................................................ 29
Figure 16: Pulse-damping Drum ................................................................................. 31
Figure 17: Installed Kistler Pressure Transducer ........................................................ 32
Figure 18: LabView Interface ..................................................................................... 34
Figure 19: LabView Block Diagram I ........................................................................ 35
Figure 20: LabView Block Diagram II ....................................................................... 36
viii
Figure 21: P-ν diagrams with Integration of Area ...................................................... 37
Figure 22: Non-dimensional Pressure vs. Volume Comparison ................................. 40
Figure 23: Cylinder Pressure vs. Crank Angle ........................................................... 41
Figure 24: P-v Diagram for SVO at Low Load .......................................................... 42
Figure 25: BHP vs. BMEP .......................................................................................... 44
Figure 26: Mechanical Efficiency vs. BMEP ............................................................. 46
Figure 27: Fuel Flow Linear Fit .................................................................................. 48
Figure 28: BSFC vs. BMEP ........................................................................................ 49
Figure 29: Brake Thermal Efficiency vs. BMEP ........................................................ 51
Figure 30: Volumetric Efficiency vs. BMEP .............................................................. 53
Figure 31: RPM vs. BMEP ......................................................................................... 54
Figure 32: Air-fuel Ratio vs. BMEP ........................................................................... 55
Figure 33: Net Exhaust Temperature vs. BMEP ........................................................ 57
Figure 34: Potential CHP Efficiency vs. BMEP ......................................................... 59
Figure 35: Exhaust Energy vs. BMEP ........................................................................ 60
Figure 36: Fuel Flow vs. BMEP ................................................................................. 61
Figure 37: U.S. Electricity Flow, 2007 (Quadrillion Btu) .......................................... 62
Figure 38: U.S. Energy Flow, 2007 (Quadrillion Btu) ............................................... 63
Figure 39: Schematic of CCHP System ...................................................................... 65
Figure 40: Turbocharger Diagram .............................................................................. 67
Figure 41: Thermocouple and Pressure Transmitter Sketch ....................................... 69
Figure 42: Compressor Inlet Attachment .................................................................... 70
Figure 43: Compressor Outlet Attachment ................................................................. 71
ix
Figure 44: Turbine Inlet Attachment .......................................................................... 72
Figure 45: Turbine Outlet Attachment ........................................................................ 73
Figure 46: Stainless Steel Fin Calculation .................................................................. 74
Figure 47: Dwyer Differential Pressure Transmitters ................................................. 75
Figure 48: Entire Turbocharger Instrumentation ........................................................ 76
Figure 49: Steam Generator ........................................................................................ 78
Figure 50: Inside View of Steam Generator ............................................................... 79
Figure 51: Coolant Heat Exchanger Before Insulated ................................................ 81
Figure 52: CCHP Monitored Systems ........................................................................ 83
Figure 53: Automated Logic Interface ........................................................................ 84
Figure 54: Mollier Diagrams for Compressor and Turbine ........................................ 87
Figure 55: Isentropic Compressor Efficiency ............................................................. 90
Figure 56: Isentropic Turbine Efficiency.................................................................... 91
Figure 57: Compressor Map ....................................................................................... 93
Figure 58: Pressure Ratio vs. Mass Flow Rate ........................................................... 94
Figure 59: Heat Transfer to Steam Generator (No. 2 Diesel Fuel) ............................. 95
Figure 60: Heat Transfer to Steam Generator (Soy Biodiesel) ................................... 96
Figure 61: Heat Transfer to Water Loop (No. 2 Diesel Fuel) ..................................... 97
Figure 62: Heat Transfer to Water Loop (Soy Biodiesel)........................................... 98
Figure 63: Air-Flow Nozzle Chart ............................................................................ 110
Figure 64: Recycled Grease Process Heat Exchanger .............................................. 111
Figure 65: Heat Exchanger Engineering Drawing .................................................... 112
x
ABSTRACT
Biodiesels are fuels that are made from renewable oils that can usually be used in
diesel engines without modification. These fuels have properties similar to fossil diesel
oils and have reduced emissions from a cleaner burn due to their higher Oxygen content.
The current and impending energy and environmental crises have revitalized the need to
find more viable renewable resources. The present work investigates the performance of
four types of diesel fuels in a 6 HP single-cylinder compression ignition (CI) engine. The
fuels of interest here are No. 2 diesel fuel, Soy biodiesel, waste vegetable oil (WVO)
biodiesel, and Canola oil (SVO). The biodiesels and vegetable oil showed higher brake
specific fuel consumption, lower brake thermal efficiency and slightly lower brake
horsepower. The Soy biodiesel was also tested against the No. 2 diesel fuel in a combined
cooling, heating and power (CCHP) system at Carnegie-Mellon University. The prime
mover for that system was a turbocharged 42 HP four-cylinder CI engine. The total
CCHP efficiency and turbocharger efficiency were monitored at various load settings.
The CCHP efficiency was 3 percent lower at full load for the Soy biodiesel compared
with No. 2 diesel fuel. The turbocharger efficiency increased significantly across all load
settings when the engine was fueled with Soy biodiesel instead of No. 2 diesel fuel. Thus
it is possible that biodiesel fuels may work more effectively than fossil diesel in certain
applications.
1
CHAPTER 1
INTRODUCTION
In 1911, Rudolf Diesel presented the world with the compression ignition engine,
which at that time did not have a specific fuel. Diesel claimed that the engine could be
fed by vegetable oils which could help the agricultural development in countries using
this engine. Biodiesels are derived from vegetable oils or animal fats, more specifically
the alkyl esters from these. The esters from vegetable oils are considered to be superior
since they have a higher energetic yield and essentially no engine modifications are
necessary for their use. Biodiesels have been traced back to the mid-1800s, where
transesterification (described in Chapter 3) was used to make soap and the alkyl esters
(biodiesels) were just considered byproducts. Early feedstocks were corn, peanut, hemp
oils, and tallow.
In 1973, the OPEC nations cut down their oil exports to the West, resulting in the
oil crisis of the 1970s. Prices of oil increased dramatically and people started to look into
renewable sources of energy. Many of the federal renewable energy programs were
initiated during the 1970s including the National Renewable Energy Laboratory. Once
prices began to fall in 1979, tax incentives and other support for the renewable energy
industry ended thanks to shortsighted policies. The initial steps taken toward a renewable
energy initiative were abandoned.
In recent years, oil prices have been rising rapidly again and there is a major
concern for the long term availability of fossil fuels. This and the growing concern for
2
our environment have created a much larger market for renewable resources. The idea of
using vegetable oils instead of fossil diesel fuels has resurfaced as a way to minimize the
net carbon footprint left by emissions from compression ignition (CI) engines. Straight
vegetable oils (SVOs) have their fair share of problems in unmodified CI engines. These
problems include: cold-weather starting; plugging and gumming of filters lines, and
injectors; engine knocking; coking of injectors on piston and head of engine; carbon
deposits on piston and head of engine; excessive engine wear; and deterioration of engine
lubricating oil. Vegetable oils decrease power output and thermal efficiency while
leaving carbon deposits inside the cylinder. Most of these problems with vegetable oil are
due to high viscosity, low cetane number, low flash point, and resulting incomplete
combustion shown in Table 1.
Table 1: Fuel Property Comparison of Vegetable Oils with No. 2 Diesel Fuel
To avoid some of these problems, vegetable oils have been converted via a
chemical process (transesterification) to result in a fuel more like fossil diesel. The
3
resulting fuel is biodiesel, a biodegradable and nontoxic renewable fuel. Furthermore,
biodiesels have reduced molecular weights (in relation to triglycerides), reduced
viscosity, and improved volatility when compared to ordinary vegetable oils (Table 2).
Most CI engines can run on biodiesels without modifications; however to optimize
combustion the injection timing should be adjusted. There may be some long term
problems that are yet to be quantified along with large scale availability and related
consequences on the agricultural sector. Overall, biodiesels have great potential and
deserve more attention and use.
Table 2: Some Fuel Properties of Various Biodiesels
1.1 Problem Statement
There has been plenty of research done so far on emissions testing and biodiesel
production. Research in the area of biodiesel has shifted towards making it more
economically feasible by lowering production costs and increasing the energetic yields
from various feedstocks. Where the research has been lacking is in relation to the better
characterization of the performance of these fuels in all possible diesel applications.
4
1.2 Scope of Present Work
The present work compares the performance of various biodiesels with fossil
diesel in multiple engine applications such as combined cooling heating and power
(CCHP) systems, turbocharging, and the use of biodiesel as a transportation fuel. The
goal of this work is to determine the usefulness of various biodiesels in a fully integrated
combined cooling heating and power (CCHP) system, as well as in a single cylinder test
engine. The CCHP performance will determine the suitability of biodiesel for stationary
diesel engine cogeneration plants. The turbocharger installed in the diesel engine for the
CCHP system will be instrumented and the turbocharger efficiency will be monitored.
The single cylinder test engine will provide data on the general performance of biodiesel
in diesel engines along with detailed P-ν (indicator) diagrams.
1.3 Outline of Thesis
This thesis will describe the experimental setups in detail, including the main
components, instrumentation, and data acquisition systems (DAQ). Chapter 3 will
provide a brief overview of the chemical processes involved in producing biodiesel. It
will also cover the specific methods used to make the biodiesel samples that were tested.
In Chapter 4, an entire engine performance analysis will be shown from the completely
instrumented single cylinder diesel engine. In Chapter 5, the thermodynamic analysis of
the CCHP system will be shown, including two heat recovery methods (steam generating
and hot water heating) as well as the enthalpy changes associated with the turbocharger.
5
CHAPTER 2
BACKGROUND
For the past decade, there has been increased interest in using biodiesels instead
of fossil diesel in CI engines. The use of biodiesel was not economically feasible until
recently due to rise in prices of fossil fuels. There is a significant amount of research
being conducted to lower the cost of producing biodiesel as well as to increase its
performance in CI engines.
Biodiesels have a higher viscosity and specific gravity than fossil diesel, which
affects fuel consumption, injection timing, and spray pattern. Since the cetane number, a
measure of combustion quality, for biodiesels is higher when compared with fossil diesel
(No. 2 fuel oil), a shorter ignition delay will result which will require an advance of
combustion timing. Canakci and Van Gerpen showed that B100 (100% soy biodiesel)
had a higher brake specific fuel consumption (BSFC) when compared to fossil diesel.
This reflects its lower heating value (about 12% lower than diesel).
6
Lin (2006) presented the following graphs (Figure 1) to compare biodiesel with
fossil diesel:
Figure 1: Fuel Consumption Rate and BSFC vs. RPM, Lin (2006)
In the above figures, the ASTM No. 2D diesel was obtained by the Chinese
Petroleum Company in Taiwan. The commercial biodiesel was provided by a large
biodiesel producer in the USA, it was produced from soybean oil and methyl alcohol via
transesterification. Sample 1 was produced by adding 50 wt% petroleum ether and 0.5
wt% distilled water to remove impurities in the coarse biodiesel. Sample 2 was produced
by putting sample 1 through a peroxidation technique to further promote the fuel
properties of sample 1. The fuel consumption rate was lowest for No. 2 diesel and highest
for commercial biodiesel. The higher fuel consumption rate and lower horsepower output
for commercial biodiesel resulted in higher brake specific fuel consumption than with
No. 2 diesel.
7
Figure 2: Brake Thermal Efficiency and Exhaust Gas Temperature vs. RPM, Lin (2006)
From Figure 2, the brake thermal efficiency is higher for the two samples of
biodiesel, but not the commercial biodiesel due to higher oxygen content (resulting in
more complete combustion). The exhaust temperatures are higher with the No. 2 diesel
compared to the commercial biodiesel. Lin (2006) showed that biodiesel has about 10
wt% of oxygen content which may allow for more complete combustion, thereby
reducing the emissions of unburned hydrocarbons (UHC) and carbon monoxide (CO).
8
Figure 3: CO, CO2, and NOx Emission Indices vs. RPM, Lin (2006)
The CO and carbon dioxide (CO2) emission indices are lower with the
commercial biodiesel but the nitrous oxide (NOx) emission index is higher with the
commercial biodiesel when compared to No. 2 diesel. Typically, biodiesel emissions in
the form of CO, particulate matter (PM), and UHC were 10-20% lower than in diesel. Lin
(2006) explained that the main concern with diesel engines and their emissions is with
nitrogen dioxide (NO2), a compound of NOx. NO2 has a strong capacity to absorb
9
infrared rays; therefore, it is a contributing factor in global warming (250 times worse
than CO2 at the same concentration). NO2 is very stable and has a high longevity and can
last approximately 150 years in the atmosphere.
Agarwal (2007) ran a series of tests with biodiesel blends. He showed that the
thermal efficiency increased slightly with higher percentages of biodiesel as shown
below:
Figure 4: Thermal Efficiency vs. BMEP, Agarwal (2007)
10
The brake specific fuel consumption (BSFC) was lower at low load with higher
percentages of biodiesel, but higher at higher loads than diesel fuel as shown in Figure 5:
Figure 5: BFSC vs. BMEP, Agarwal (2007)
Also, the exhaust temperature increased with increasing percentages of biodiesel,
as shown in the following graph:
Figure 6: Exhaust Temperature vs. BMEP, Agarwal (2007)
11
One type of biodiesel can be produced from restaurant grease. In 1990,
approximately 1.1 billion kilograms of waste grease was collected from restaurants and
fast-food establishments in the United States. Canakci (2001) showed that the best results
of waste vegetable oils without transesterification came from 10-20% blends in indirectinjection engines. Deposits in the cylinder increased but did not appear to affect
performance. If the waste oils were put through transesterification, it seems all
specification values can be met except for the cold filter plugging point (mostly over -8
°C). Used frying oil emissions contained slightly lower amounts of HC, CO, and
particulate matter but increased NOx. Also, acid-catalyzed frying oil had low viscosity
while base-catalyzed frying oil had high viscosity. The viscosity results correlate to the
percentage of ester yield, indicating that some fuels probably had substantial amounts of
unreacted and partially reacted oil since acid-catalysts improve reaction better than basecatalysts.
Biodiesel also proves to serve as a better lubricant than fossil diesel due to the
lower deterioration in density through the engine operation. This could lower the wear on
vital engine components, lower fuel dilution and moisture content. Though the long term
use of biodiesels and any associated derivatives are not well known, biodiesels appear to
be compatible with CI engines and result in decreased emissions. If the cost of biofuels is
managed properly, their use is bound to increase.
12
CHAPTER 3
BIODIESELS
A brief overview of the production of biodiesels and the use of various
components used is given here. There are many methods that can be used to make
biodiesels with multiple combinations of catalysts, neutralizers, and feedstocks. After
reviewing the general and most popular methods, the ones used to create our test samples
will be explained in greater detail.
3.1 Overview
3.1.1 Organic Chemistry
The major components of vegetable oils are triglycerides. Triglycerides are esters
of glycerol with long-chain acids (fatty acids). The composition of vegetable oils varies
with the plant source. The fatty acid profile describes the specific nature of fatty acids
occurring in fats and oils. The chemical and physical properties of fats and oils and the
esters derived from them vary with the fatty acid profile.
Transesterification is the process where an alcohol and an ester react to form a
different alcohol and a different ester. For biodiesel, an ethyl ester reacts with methanol
to form a methyl ester and ethanol. These ethyl esters react with methanol to form
biodiesel and glycerol. As mentioned above, the purpose of transesterification is to
reduce the viscosity of the oil so that it has properties closer to that of regular diesel used
in CI engines. Methanol is the preferred alcohol for obtaining biodiesel because it has the
13
lowest cost and it is readily available. However, for the reaction to occur in a reasonable
time, a catalyst must be added to the mixture of the vegetable oil and methanol to
accelerate the speed of a reaction. Below is a figure showing the endothermic (requiring
heat) chemical reaction behind the transesterification process.
Figure 7: Transesterification Reaction
R represents a mixture of various fatty acid chains and therefore must be defined
based on the oils in use. The subscript 3 indicates the number of moles needed to satisfy
the formation of the methyl esters. This model only states the molar ratios of starting
materials and products however; the molar ratios may need to be varied to obtain a more
complete reaction. Typically, 6 moles of alcohol are used for every mole of triglyceride
so that the reaction proceeds in the „forward‟ direction. Not all reactions complete
quickly and for some it takes considerable time before the starting materials and reaction
products are present in constant amounts, indicating equilibrium has been reached. A
reaction can also occur in the reverse direction (from right to left), so to force the
equilibrium in the direction of the desired products, one or more parameters of the
14
reaction may need to be changed. These include the molar ratio, temperature, pressure
and use of a catalyst.
3.1.2 Feedstock
Feedstocks for production of biodiesels are vegetable oils (soybean, canola, palm,
and rapeseed), animal fats (beef, tallow, lard, poultry fat, fish oils) or recycled grease
(mix of the above two). All of the above feedstocks contain triglycerides, free fatty acids
(FFAs) and other contaminants. The proportions vary in level depending on the feedstock
and these variables affect the chemical reactions needed to transform the primary raw
materials (feedstock and alcohol) to create the biodiesels. Commercially available
vegetable oils are made up of a small percentage of FFAs, but crude vegetable oil may
contain more FFAs and phospholipids, which are removed in two processes: refining and
degumming, respectively. The technology required depends on whether the vegetable oils
are refined, degummed, or crude. Animal fats and recycled grease have high levels (up to
15% concentration) of FFAs. The FFA content affects the process and yield associated
with the final product, and thus these feedstocks (greater that 1% concentration) must be
pretreated before the reaction can begin. Other contaminants also affect the feedstock
preparation necessary before it can be used in the reaction.
3.1.3 Alcohol
In order to form the biodiesel, a primary alcohol is coupled with the feedstock to
form the esters. The most common alcohol is methanol but ethanol, isopropanol and butyl
(derived from butane) can also be used. The key quality parameter associated with the
15
process of transesterification is the water content. If the water content is high, it results in
low yields, high levels of soap, and leftover FFAs/triglycerides. Therefore, it is important
to use stronger alcohols, since the weaker ones are hygroscopic (absorb water from the
air). Some alcohols also require higher operating temperatures, longer mixing times, and
lower mixing speeds which result in higher operation costs and lower throughput. The
decision of which alcohol to use with each process is determined by cost, amount needed,
and ease of recycling. Quality requirements on fuel and water content in ethanol make
methanol the most popular choice among alcohols. Conventionally, methanol is not
renewable since it is normally generated from either natural gas or coal gas and steam
however methanol can now be produced from renewable biomass materials (like wood,
and black liquor from pulp and paper mills).
3.1.4 Catalyst
Catalysts are required to promote an increase in solubility which allows the
reaction to occur at a faster rate. These include base, acid or enzyme catalysts. The most
common catalysts are sodium hydroxide (NaOH) and potassium hydroxide (KOH). Most
base catalyst systems use vegetable oil as the feedstock, but base catalysts are highly
hygroscopic, so absorption of chemical water occurs when the catalyst is dissolved in the
alcohol. This leads to poor quality biodiesel that may not meet the ASTM standard.
Typically, base catalysts are used for vegetable oil processing plants because FFA and
water content is low and the reaction is fast.
16
3.1.5 Neutralizer
A neutralizer is used to remove the base or acid catalyst from the
biodiesel/glycerol products. If using a base catalyst, the neutralizer must be acidic, and if
using an acid catalyst, the neutralizer must be a base. Neutralization may occur when the
base catalyst is added to convert the remaining triglycerides, or if the biodiesel is being
washed, the base catalyst may be added to the wash water. Hydrochloric acid is typically
used as the acid neutralizer because it is cheap, but one advantage of using phosphoric
acid is that the resulting salt may be sold as a chemical fertilizer.
3.1.6 Glycerol
Glycerol is not a component in making biodiesel but it is the primary byproduct of
the reaction. In trying to make biodiesel economically feasible in today‟s market, it is
extremely important to consider glycerol‟s profitability since it has many marketable
applications. Glycerol in its pure form is a sweet-tasting, clear, colorless, odorless,
viscous liquid. It is completely soluble in water and alcohol. Potential customers include
pharmaceutical and cosmetic manufacturers, and many others. It can be used as a solvent,
sweetener or in the manufacturing of dynamite, cosmetics, liquid soaps, candy, liquors,
inks, and lubricants. Usually about 10% of the resulting mixture from the reaction can be
recovered as glycerol.
17
3.2 Waste Vegetable Oil (WVO)
The biodiesel used in the present work is derived from waste vegetable oil (WVO).
It has been made locally by a Lehigh University staff member (Mr. Jason Slipp). These
biodiesels are made from waste cooking oil supplied by various restaurants, and thus FFA
content of each batch is slightly different. Since the waste vegetable oil is collected free
of charge (except transport), a gallon of biodiesel can be made for less than $2.00.
Currently, it is made in 50 gallon batches at a frequency of about once every week. If
there was a higher demand for his biodiesel it could be made even more cheaply. The
batch used in our tests was collected from restaurants that use peanut, soy and canola oil.
The restaurants are local so the transportation cost associated with collection of the waste
vegetable oil is insignificant unless produced on a mass scale. The methanol used in the
process is bought in 55 gallon drums for about $200, about 10 to 13 percent of which can
be recovered during the process.
The waste vegetable oil is filtered using a colander before entering the feedstock
tank. It is then filtered again while pumped to the reactor vessel. While the 50 gallons of
feedstock is pumped into the reactor, the tank is heated to about 130 °F. The methanol
and NaOH is put into a separate tank (10 gallons of methanol and an undetermined
amount of NaOH). To determine the amount of NaOH needed to drive the reaction, a
sample of feedstock and methanol is tested with Lye water (the current batch required
852 grams of NaOH for 50 gallons of waste vegetable oil). This step is known as
18
titration, the quantitative chemical analysis used to determine the unknown concentration
of a reactant. A photograph of the process is shown in Figure 8.
Figure 8: Titration
The Lye water (NaOH and water) is added using a liquid dropper to the feedstock and
methanol mixture until the color turns magenta. The amount of Lye water is then
recorded and a calculation is made for the amount of NaOH needed for the particular
batch of biodiesel. The feedstock is heated for two hours to remove all of the water and
then the NaOH is added to the methanol in the mixing tank and then the mixture is slowly
added to the reactor tank. The feedstock, methanol and NaOH are circulated in the reactor
for another two hours. The mixture is allowed to settle for 48 hours and then the glycerin
19
is separated by pumping from the bottom of the tank into the glycerin tank. Once all of
the glycerin is out of the bottom (which can be seen as the fluid in the clear hose changes
over to biodiesel), the solution (now biodiesel) is pumped to the wash tank where it is
washed 3 times (20 gallons of water for each rinse for the 50 gallons of feedstock). The
Mono/Di glyceride molecules are water soluble so they bind with the water molecules
and create a soapy mix at the bottom of the wash tank. Now the biodiesel is ready for use
and is pumped into the biodiesel storage tank. A photograph of the mixture at the various
stages throughout the process is shown in Figure 9:
Figure 9: Biodiesel Mixture throughout Entire Process
From http://www.ballarat.edu.au/projects/ensus/case_studies/biodiesel/
20
A photograph of the system used to make the biodiesel from waste vegetable oil
described here is shown in Figure 10:
Figure 10: Recycled Grease Manufacturing System
21
Figure 11 shows a schematic of the entire system used to make the waste
vegetable oil (WVO) biodiesel from start to finish:
Figure 11: Schematic of Recycled Grease Manufacturing System
22
3.3 Soy Biodiesel
Soy biodiesel is typically made by the transesterification process. The soybeans are
put through an acid correction process and then transferred to a principal reactor. The
catalyst (KOH) and alcohol (methanol) are homogenized and then transferred to the
principal reactor. The reaction takes about 40 minutes at temperatures between 30 and 40
°C. The biodiesel is separated from the raw glycerin continuously using a centrifuge. The
raw glycerin can be purified further, and then sold to market.
The soy biodiesel used in study was produced by CTI Biofuels in Pittsburgh, PA.
The heating value is 37,257 kJ/L (18,087 Btu/lb m). This biodiesel was also tested via
ASTM methods; the results are shown in Table 3.
Table 3: Soy Biodiesel Properties
Analysis
Method
Results
Density, kg/m3 @ 15 °C
ASTM D4052
885.6
Kinematic Viscosity (cSt) ASTM D445
4.08
@ 40 °C
Cetane Number
ASTM D613
23
49.6
CHAPTER 4
SINGLE-CYLINDER DIESEL ENGINE
The overall effectiveness of the various fuels in a 6 HP single-cylinder CI test
engine is described here. The fuels to be tested in this engine are: soy biodiesel, WVO
biodiesel, Canola Oil (SVO), and No. 2 fuel oil (fossil diesel). The biodiesel samples will
be 100% biodiesel (B100) and the straight vegetable oil (SVO) will also be 100%. The
engine with a hydraulic dynamometer is used to measure the power output, specific fuel
consumption, P-ν characteristics, etc. for these fuels to provide an accurate means of
comparing their performances.
4.1 Experimental Setup
4.1.1 Engine
The engine that was used for the fuel tests was made by Yanmar Co., LTD. The
engine specifications are shown in the table below:
Table 4: Yanmar Engine Specifications
Engine Model:
L70V
High Idling
3800+/-30
Engine Type:
4-stroke, Vertical
Combustion
Direct Injection
Cylinder, Air-
System
Cooled Diesel
No. of Cylinders
1
Cooling System
Forced Air by
Flywheel Fan
24
Bore x Stroke
Displacement
3.07 x 2.64 in
19.5 cu. in
Lubricating
Forced Lubrication
System
with Trochoid Pump
Dimensions (L x W
14.9 x 16.6 x 17.8 in
x H)
Continuous Rated
3600 RPM
Maximum Rated
3600 RPM
Output
5.8 hp SAE
Output
6.4 hp SAE
4.3 kW
4.8 kW
The engine has been instrumented with an in-cylinder pressure transducer, a
magnetic pickup/resolver for crank angle, thermocouples for intake and exhaust gases,
load cells and tachometer for the dynamometer, a weighted fuel tank, and an ASME
nozzle for the air flow measurement. With these instruments P-ν diagrams, Brake
Specific Fuel Consumption (BSFC), Exhaust Temperature, Brake Horsepower (BHP),
Mechanical Efficiency, Thermal Efficiency, and Volumetric Efficiency for varied loads
could be determined. The engine setup is shown in Figure 12 on the next page.
25
Figure 12: Single-Cylinder Test Setup
26
4.1.2 Load Cells
Two load cells, SM S-Type made by Interface, were used to monitor the fuel
weight and the engine load. They are accurate to within 0.03% full scale. A close-up
picture of the load cell is shown in Figure 13:
Figure 13: Load Cell
4.1.3 Dynamometer
The dynamometer (DY-7D from Go-Power Corporation) is used to load the
engine and measure the shaft rotary power (horsepower) by measuring the torque and
RPM of the engine. The schematic in Figure 14 shows the hydraulic dynamometer used
to create load on the engine. There are hoses for the water-in, water-out and the one-way
valve to allow air to re-enter the absorption unit when the flow is decreased. In the
present setup, a clamp to pinch the hose was used at high load instead of the one-way air
valve to prevent water from going out through the air hose.
27
Figure 14: Dynamometer Load Control System
The maximum torque capacity of the absorption unit at a given speed is reached
when the dynamometer is completely filled with water. When this occurs, the torque
increases in direct proportion to speed at 4.5 ft-lbf / 1,000RPM. The maximum power
capacity for this absorption unit at 3,750 RPM (our operating range) is about 11 HP,
although the engine is rated for 6 HP. At full throttle, the average torque system errors are
+/- 0.1 lbf, or +/- 0.6%.
The speed of the engine is measured by a mechanical tachometer which is driven
off of the end of the device shaft. It is a centrifugal type tachometer and uses an internal
three ball governor movement to actuate the tachometer pointer. The true RPM reading is
within 2% at full throttle and even better at lower values.
28
4.1.4 Resolver
A resolver and magnetic pickup were used to determine the volume of the
cylinder throughout the cycle. The resolver is an electromechanical device that converts
angular displacement values into digital values. As the drive shaft changes its angular
position, the resolver sends an analog signal to an analog to digital converter (A-D
Board) determining the location of the piston in the cylinder. The resolver can only work
with a reference point, which is determined by the magnetic pickup (Hall Effect sensor).
In this setup,(see Figure 15), the timing mark is a screw installed on an iron disk attached
to the drive shaft and enables a fixed point to correspond to top dead center (TDC).
Figure 15: Setup for Magnetic Pickup
29
4.1.5 Manometer and Pulse-damping Drum
The air flow is measured by drawing the intake air through a precision long-radius
ASME flow nozzle (0.75 inches in diameter, Figure 16) into a pulse-damping drum and
then out through a flexible air hose to the engine intake. All of the air that enters the
engine is drawn through the nozzle. Measuring the pressure difference across the flow
nozzle allows the calculation of the air flow to within a high degree of accuracy. The
pressure difference is measured in inches of water using a manometer. The pulsedamping drum is connected via tygon tubing to the manometer.
30
Figure 16: Pulse-damping Drum
31
4.1.6 Pressure Transducer
The pressure transducer installed in the engine cylinder is a Kistler Type 701A
(Figure 17). This piezoelectric transducer was retrofitted with a Type 7505 water cooler
to minimize the effect of high temperatures. The charge signal of the transducer (in picoCoulombs) is amplified and transformed to a proportional output voltage in the Kistler
charge amplifier. The maximum range is 1,500 psi with a sensitivity of 2.867mV/psi.
Between 100 and 350 °C, the error in pressure measurement can be no higher than +/3.5%.
Figure 17: Installed Kistler Pressure Transducer
32
4.1.7 Thermocouples
The thermocouples used were miniature K-type from Omega Engineering, Inc.
They are accurate to within 2.2 °C or 0.75% (whichever is greater) under the conditions
used here. One was installed to monitor the temperature of the incoming air and the other
for the exhaust gases in the exhaust manifold.
4.2 Procedure and DAQ
The engine throttle and dynamometer load settings were adjusted manually,
making it difficult to repeat exact settings. Therefore, it was decided to cycle through
various load settings (from the dynamometer) at constant throttle settings. For each
throttle setting, 4-5 different loads were taken allowing us to analyze the data and
interpolate it for each fuel type. All data was acquired after the load was adjusted and the
engine reached steady state conditions. The process lasted about 25 to 30 minutes for
each setting.
Figure 18 shows the LabView VI used for DAQ. It displays Instantaneous Load
and RPM, Fuel Weight, Exhaust/Intake Temperatures, Cylinder Pressure Amplitude and
records all data along with a Pressure vs. Volume curve. The VI evaluated values of
Work (lbf-in), Indicated and Brake Horsepower, Indicated and Brake Mean Effective
Pressure (psig), Mechanical Efficiency, and Maximum Pressure (psig) from the P-ν
diagram.
33
Figure 18: LabView Interface
The following two figures are the block diagrams (left-side and right-side) for the
LabView VI file. The “while loop” is set to take data every 5 seconds and interpolate the
P-ν diagram and to calculate and record the data mentioned above.
34
Figure 19: LabView Block Diagram I
35
Figure 20: LabView Block Diagram II
4.3 Results
4.3.1 P-ν diagrams
The performance of an engine depends on its thermodynamic cycle, and the
pressure and volume variations in the cylinder. The pressure data for the gas in the
cylinder over the operating cycle of the engine can be used to calculate the net work
transfer from the gas to the piston. The indicated work per cycle is obtained by the cyclic
integration around the area enclosed by the P-ν diagram:
Wc,i
Pdv
36
In order to analyze the P-ν diagrams for all four fuels (soy, WVO, SVO, diesel),
the engine was run at wide open throttle, with various load settings. The P-ν diagrams
shown below indicate the areas enclosed and the load setting for each one of the four
fuels:
Figure 21: P-ν diagrams with Integration of Area
The maximum pressure varies for each of the fuels; No. 2 diesel is 718 psig, soy
biodiesel is 739 psig, WVO biodiesel is 756 psig, and SVO is 750 psig. From these areas,
37
the indicated power per cycle (IHP) and indicated mean effective pressure (IMEP) were
calculated for each fuel. The IHP was calculated using the following formula:
IHP (hp)
Wc ,i N
( A)
nR
ft lbf
N power _ strokes
( )
power _ stroke 2
min
ft lbf
(33,000)
HP min
where nR is the number of crank revolutions for each power stroke per cylinder. For fourstroke engines, this value is 2. Indicated power differs from the brake power by the power
absorbed in overcoming engine friction, driving accessories and the pumping power. The
Indicated Mean Effective Pressure (IMEP) was calculated using this formula:
IMEP ( psig)
Wc , i
Vd
ft lbf
in
(12)
HP min
ft
N power _ strokes
( )
(Vd )in 3
2
min
IHP (33,000)
where Vd is the displaced or swept volume of the cylinder, 19.5 in3 for this engine. One
set of the calculated IHP and IMEP values for each fuel are shown in the following table:
Table 5: Calculated Values from PV-diagrams
Fuel Type
RPM
Wc,i
IHP
IMEP
+/- 2%
+/- 0.6%
+/- 0.6%
+/- 0.6%
No. 2 Diesel
3634
2510
11.52
128.7
Soy
3645
2500
11.51
128.2
WVO
3504
2360
10.44
121.0
SVO
3636
2450
11.25
125.6
38
Notice that the RPM for the WVO is much lower than the RPM for all other fuels
in the Table 5; this indicates that the engine was slowing down to create the torque
required to rotate the shaft linked to the dynamometer. Thus, the engine is capable of
undergoing slightly lesser loading when fueled with the WVO. This is due to the lower
heating value of the recycled grease producing lower horsepower in the engine.
Considering that these P-ν diagrams were taken at slightly different loads, the area
enclosed by the diagrams must be normalized. This was done by multiplying the
pressures by the highest IMEP (from No. 2 diesel fuel) then dividing by the IMEP for the
specific fuel. Normalizing the graphs this way expands the area enclosed with the slightly
lower loads (WVO, SVO, soy biodiesel) to the area enclosed by the No. 2 diesel fuel.
Then, the pressures were non-dimensionalized by dividing all pressures by the maximum
pressure during the cycle. Doing this, all three diagrams can be plotted on the same graph
and the pressure range goes from 0 to 1. The plot with all three P-ν diagrams is shown in
Figure 22.
39
Figure 22: Non-dimensional Pressure vs. Volume Comparison
It can be seen from the normalized P-ν diagrams that the area enclosed for soy
biodiesel is slightly less than diesel since the compression strokes are almost the exact
same but the soy biodiesel expansion stroke is offset from the one with No. 2 diesel. For
the SVO, the area enclosed is less than No. 2 diesel and soy biodiesel for the same reason
previously mentioned. The same result is seen with the WVO which has less area
enclosed than all other fuels. These results are due to the heating values for each fuel (No.
2 diesel having the highest and the WVO having the lowest).
40
Figure 23 shows the cylinder pressure throughout the compression and expansion
strokes for the different fuels.
Cylinder Pressure vs. Crank Angle
750
650
SOC
EOC
550
Pressure (psig)
SOI
450
Diesel
WVO
350
Soy
SVO
250
150
EVO
IVC
50
-180
-50
-135
-90
-45
0
45
90
135
180
Crank Angle (degrees)
Figure 23: Cylinder Pressure vs. Crank Angle
Intake Valve Closed (IVC), Start of Injection (SOI), Start of Combustion (SOC),
End of Combustion (EOC) and Exhaust Valve Opened (EVO) are labeled on the graph.
These locations are based on the typical values found in Heywood (1988). The pressures
are very similar with each fuel but the WVO and SVO have higher maximum pressures
compared to the soy biodiesel and No. 2 diesel fuel.
At low load (40 psig BMEP or less), the SVO P-ν diagrams are extremely
inconsistent. The area enclosed by the diagrams fluctuate at each setting, therefore the
41
engine is producing an inconsistent amount of work. This is shown in Figure 24 for the
SVO at 5.54 lbf.
Figure 24: P-v Diagram for SVO at Low Load
Notice that during blow down (exhaust valve open) the pressure rises and then during the
intake stroke (intake valve open) the pressure also rises. The exhaust and intake strokes
should be at constant pressure throughout the stroke, slightly above atmospheric and
42
slightly below, respectively. At low load the engine temperature is low; therefore the
viscosity of the fuel significantly affects the fuel injector pump and its ability to supply
enough fuel to the cylinder. The high viscosity of the fuel can result in injection pump
seizures, and poor injection nozzle spray atomization. Poor atomization results in reduced
combustion efficiency generating high emissions of unburned hydrocarbons which can
inhibit the exhaust gases exiting the engine cylinder (demonstrated in Figure 24 above).
4.3.2 Brake Mean Effective Pressure (BMEP)
The brake mean effective pressure is an important concept for comparing
different fuels. It is the average pressure the engine can exert on the piston through one
complete operating cycle. It is the average pressure of the gas inside the engine cylinder
based on net power. BMEP is important because it is independent of the RPM and size of
the engine. Thus, all of the following plots in this chapter will be plotted against BMEP.
It is calculated the same way as the IMEP (shown above) except using the BHP (from the
dynamometer) instead of the IHP.
4.3.3 Brake Horsepower (BHP)
The brake horsepower is the actual shaft horsepower and is measured by the
dynamometer. This is the useful power of the engine and is shown for these fuels in
Figure 25.
43
Figure 25: BHP vs. BMEP
The BHP is exactly the same for the soy-based biodiesel, SVO and No. 2 diesel
fuel at full throttle, but it drops off slightly at higher load for the WVO. This agrees with
what was found previously from the P-ν diagrams in that the WVO does not produce the
same amount of power as the other fuels at extreme engine load.
44
4.3.4 Mechanical Efficiency
Part of the indicated work per cycle is used to expel exhaust gases, induct fresh
air, and also to overcome friction of the bearings, pistons, and other mechanical
equipment in the engine. The mechanical efficiency is the measure of the ability of the
engine to overcome the frictional power loss. It is calculated as follows:
m
BHP
IHP
A graph of the mechanical efficiencies associated with each of the three fuels is shown in
Figure 26 on the next page.
45
Figure 26: Mechanical Efficiency vs. BMEP
The mechanical efficiency is similar for each fuel; however the efficiency of the
soy biodiesel is higher than the other fuels across all engine loads. The mechanical
efficiency of the engine is lowest when fueled by the SVO. At full load, all fuels
converge to about 52% mechanical efficiency.
46
4.3.5 Brake Specific Fuel Consumption (BSFC)
The brake specific fuel consumption is defined as the fuel flow rate per unit
power output. It is a measure of the efficiency of the engine in using the fuel supplied to
produce work. It is desirable to obtain a lower value of BSFC meaning that the engine
used less fuel to produce the same amount of work. This is one of the most important
parameters to compare when testing various fuels. It is calculated as follows:
lbs
hr
( BHP )hp
(m f )
BSFC
 f is the mass flow rate of the fuel and BHP is the brake horsepower. BHP is read
where m
by the dynamometer, and the fuel flow rate is determined by the load cell attached to the
fuel tank. The fuel flow was measured with weight and time measurements. The fuel tank
weight was correlated with the load cell data. Using a linear fit, the fuel flow in lbs/hr
was determined. The vibrating of the engine made this difficult, thus fuel usage was
determined over a period of 25-30 minutes with the data being taken every 5 seconds.
The linear fit was very accurate for this time frame as shown in Figure 27. Much of the
disparity from the linear fit was due to engine vibrations and its effect on the load cell
that is measuring the fuel tank weight.
47
Figure 27: Fuel Flow Linear Fit
The error associated with this graph seems to be about +/- 0.1 lbs/hr. The plot for
the BSFC of the various fuels is illustrated in Figure 28.
48
Figure 28: BSFC vs. BMEP
As expected, the No. 2 diesel fuel has the best (lowest) BSFC for the conditions
used here. This is due to the higher heating value in fossil diesel compared to the other
fuels thus allowing for less fuel to be used to generate the same amount of power. The
WVO performed next best at high load. The SVO has a low BSFC at low load, but it rises
at high load.
49
4.3.6 Brake Thermal Efficiency
The thermal efficiency is the ratio of the thermal power available in the fuel to the
power the engine delivers to the crankshaft. This greatly depends on the manner in which
the energy is converted since the efficiency is normalized with the fuel heating value. The
heating value of No. 2 diesel fuel is 19,524 Btu/lb, for soy biodiesel it is 18,087 Btu/lb,
for SVO it is 17,244 Btu/lb and for WVO it is 16,885 Btu/lb. Using these values the
thermal efficiency can be determined using this formula:
( BHP )
t
(QHV )
BTU
lbm
BTU
hr
(m f )
lbm
hr
where QHV is the heating value. The thermal efficiencies for the fuels are shown in Figure
29.
50
Figure 29: Brake Thermal Efficiency vs. BMEP
The WVO has the highest thermal efficiency at high load. The No. 2 diesel fuel
performed best all-around and the soy biodiesel performed the worst compared to all
other fuels.
51
4.3.7 Volumetric Efficiency
The volumetric efficiency measures the effectiveness of an engine‟s induction
process, which consists of the air filter, intake manifold, intake port, intake valve etc. It is
only used with four-stroke engines which have distinct induction processes. It is the ratio
of the air drawn into the engine cylinder to the maximum amount of air which could have
been drawn into the engine cylinder as shown below:
v
2m a
a ,i V d N
2 (m a )
(4.33 10 5 )
lb
in 3
lb
hr
(
(V d )in 3
1 hr
)
60 min
( RPM )
rotations
min
Since engines run at high speeds, the cylinder rarely fills up to 100% of its capacity. The
volumetric efficiencies for the three fuels are shown in Figure 30.
52
Figure 30: Volumetric Efficiency vs. BMEP
The SVO has the highest volumetric efficiency out of the four fuels. The
volumetric efficiency when fueled with WVO seems to drop linearly at higher loads even
though the other graphs drop exponentially. This is because the volumetric efficiency is
inversely proportional to the RPM and at higher loads the engine RPM decreases when
fueled with the WVO (shown in Figure 31).
53
Figure 31: RPM vs. BMEP
Notice the severe drop off of RPM as the load increases in the engine. The engine
load (torque) is proportional to the horsepower divided by the RPM of the engine. Since
the WVO produces less horsepower at higher loads (25 psig BMEP and higher), the RPM
of the engine must drop to generate the same torque.
54
Figure 32: Air-fuel Ratio vs. BMEP
No. 2 diesel fuel has the best volumetric efficiency since it uses much more air
per lb of fuel, as shown by the higher air-fuel ratio. The soy biodiesel and WVO are
similar across all load settings but the SVO has a high air-fuel ratio at low load and then
joins the others at high load.
55
4.3.8 Net Exhaust Temperature
Exhaust gases of an internal combustion engine contain significant enthalpy and
may contain unburned combustion products (hydrocarbons). When the air-fuel ratio is
high, the amount of incomplete combustion products is likely to be low and when the airfuel ratio is low, there is an insufficient amount of oxygen to complete combustion. The
exhaust temperature is related to the determination of system efficiency. The net exhaust
temperatures (exhaust minus intake temperature) for these fuels are shown in Figure 33
on the next page.
56
Figure 33: Net Exhaust Temperature vs. BMEP
The net exhaust temperature is slightly higher with the soy biodiesel when
compared to the No. 2 diesel fuel and SVO. At low load the WVO exhaust temperature
was similar to the other fuels but at higher load the engine was not running at the same
BHP, therefore the exhaust temperature decreased as the load increased.
57
4.3.9 Potential CHP Efficiency
The combined heat and power (CHP) efficiency is relevant here because we are
interested in seeing the amount of recoverable heat available with this engine and with
these three fuels. This efficiency is the energy produced by the engine (horsepower and
exhaust stream) divided by the energy put in (fuel in). In an engine, typically 1/3 of the
power goes towards shaft power (BHP), 1/3 leaves through the exhaust and 1/3 is taken
away through engine cooling. In this case, the engine is air cooled therefore the heat
taken away from the engine through cooling is not measurable. The efficiency is
calculated as follows:
chp
E in
E out
m exhaust
cp
m f
T
BHP 2544
Q HV
where cp is the specific heat of the exhaust (assumed to be mostly air), and m exhaust is the
mass flow rate of the air plus the mass flow rate of the fuel. Figure 34 is a graph of
potential CHP efficiency with these fuels.
58
Figure 34: Potential CHP Efficiency vs. BMEP
The CHP efficiency is lowest with the soy biodiesel at all loads and consistently
highest with No. 2 diesel across most loads. For the WVO, the efficiency reaches
extremely high values at high load. To fully understand the CHP potential for each fuel,
the energy leaving through the exhaust gases should be viewed independently, shown in
Figure 35.
59
Figure 35: Exhaust Energy vs. BMEP
The WVO has the least amount of energy leaving the engine through the exhaust
over all load settings. At high load, the SVO has the highest amount of energy leaving
with the exhaust gases, this is due to the high mass flow rates of air and fuel resulting in a
much higher flow rate of the exhaust gases. The higher mass flow rate of fuel for the
SVO is shown in Figure 36.
60
Figure 36: Fuel Flow vs. BMEP
The fuel flow is consistently higher with the soy biodiesel than the other fuels
except for SVO at high load. At low load, the fuel flow is low for the SVO as a result of
the high viscosity of the fuel at low temperature. The diesel uses much less fuel across all
loads. The WVO fuel flow does not increase like the other fuels at high load, therefore
the engine is running lean which explains the drop in horsepower.
61
CHAPTER 5
CCHP TESTS WITH A FOUR-CYLINDER DIESEL ENGINE
A four-cylinder diesel engine at Carnegie-Mellon University was used as the
prime mover for a combined cooling, heating and power (CCHP) system. CCHP systems
can yield better overall efficiency of a system. Typically, CCHP efficiencies range
between 65% and 80% while the typical U.S. power plant efficiency is around 31%. This
is shown in the U.S. electricity flow diagram (Figure 37):
Figure 37: U.S. Electricity Flow, 2007 (Quadrillion Btu)
In a CCHP system, the transmission and distribution losses (T & D losses) can be
eliminated since most of the electricity generated is used on site (some electricity may be
sent to a nearby user, but the distance will be negligible). The conversion losses in the
above diagram can be minimized since most of that is in the form of waste heat that can
62
be recovered. The recovered heat can be used to heat domestic hot water, space heat,
generate steam, etc. In this CCHP system, a biodiesel-fueled engine is used to generate
electricity, a steam generator to recover the heat from the exhaust gases, and a heat
exchanger to recover the heat in the engine coolant. Soy-based biodiesel and fossil diesel
will be tested to compare their performances while monitoring the entire system.
The predicted shortage of fossil fuels, and the fact that more than half of the U.S.
energy supply comes from fossil fuels (Figure 38) exemplifies the need to find viable
substitutes for fossil fuels and petroleum products in particular. Biodiesels are renewable
fuels that can help the U.S. minimize its net carbon footprint while decreasing the
dependence on foreign oil. The goal of this project is to determine the feasibility of using
biodiesel in a CCHP application.
Figure 38: U.S. Energy Flow, 2007 (Quadrillion Btu)
63
5.1 Experimental Setup
The four-cylinder diesel engine is part of the Intelligent Workplace Energy Supply
System (IWESS) located at Carnegie-Mellon University (CMU). The IWESS is a lived
in, and researched office space consisting of solar receivers, parabolic trough reflectors,
two absorption chillers, a biodiesel engine generator with heat recovery, convective fan
coils, radiant mullion pipes, and ceiling panels. This integrated CCHP system also has a
ventilation unit including enthalpy recovery, an air based heat pump, and air
humidification by a solid desiccant. The objective of IWESS is to provide a healthy,
productive, and comfortable environment for its occupants, consisting of graduate
students and faculty members of the architecture and mechanical engineering
departments. The students and faculty have been involved in the design, installation, test,
and evaluation of the entire system. The purpose of this research is to determine the
effectiveness of biodiesel in the engine generator along with its heat recovery equipment.
Therefore, the rest of the IWESS system will not be included in this analysis. A
schematic for the system is shown in Figure 39.
64
Figure 39: Schematic of CCHP System
During the heating season, the electrical power generated is fed into the campus
grid; the coolant energy and steam provide space heating for the IWESS. During the
cooling season, the electrical power is still fed into the campus grid but the coolant
energy is routed to the solid desiccant dehumidification system to regenerate solid
desiccant. The steam (at 87 psig) is directed to the steam driven double effect absorption
chiller.
5.1.1 Engine
The diesel engine used for this system is a 42 horsepower John Deere
(4024TF270) turbocharged four cylinder engine. The fuel lines and gaskets have been
modified by the manufacturer to handle 100% biodiesel fuel (B100).
65
The engine specifications are shown in the following table:
Table 6: Specifications for John Deere 4024TF270
Model
4024TF270
Aspiration
Turbocharged
Number of Cylinders
4
Length—in. (mm)
26.1 (662)
Displacement—L (in3)
2.4 (149)
Width—in (mm)
22.3 (566)
Bore and Stroke—in.
3.4 x 4.1 (86 x 105)
Height—in. (mm)
30.4 (772)
Compression Ratio
20.5 : 1
Weight, dry—lb (kg)
553 (251)
Engine Type
In-line, 4-Cycle
(mm)
This engine is fully instrumented to complete a thermodynamic analysis of the
engine. The fuel system has a differential flow meter that measures the flow of fuel going
into the engine and the flow returning from the engine to provide a differential output.
The fuel temperature is measured by a strap-on thermocouple since the fuel flow lines are
only ½″ in diameter and an in-flow thermocouple would obstruct the flow. An ultrasonic
level sensor was installed at the top of each 300 gallon fuel tank to provide a back fuel
consumption measurement.
The air intake has an inline flow meter and thermocouple installed. The exhaust
has three temperature measurements, and an emissions suite. The exhaust temperatures
are measured at the engine exhaust, at the exhaust of the steam generator, and at the
stack. The steam side of the exhaust system has two pressure measurements and one flow
measurement.
66
5.1.2 Turbocharger
The engine is turbocharged, making the turbocharger performance another focus
of research. A turbocharger uses the pressure and flow rate of the exhaust gases to power
a turbine which is linked by a shaft to a compressor (Figure 40). The compressor
increases the pressure of the intake air so that more oxygen per unit volume can fit into
the engine cylinders allowing more fuel to fit, hence increasing the power output of the
engine.
Figure 40: Turbocharger Diagram
From http://www.kickflop.net/wp/myimages/turbo-diagram.jpg
Air is drawn to the centrifugal compressor impeller because of the low pressure
created in front of the impeller by the action of the impeller itself. The flow changes from
the axial direction to the radial direction in the compressor impeller significantly
increasing the tangential momentum of the fluid (air).
67
This is due to the work input dictated by the Euler turbomachinery equation
shown below.
Wx
U 2C
2
U1C
1
where Wx is the work delivered to the flow per unit mass flow rate m, giving a
temperature and pressure rise. U is the blade speed and Cθ is the absolute velocity in the
tangential direction. The work goes into changing the velocity leaving compressor with
respect to the velocity entering the compressor, it is affected by the angle changes (from
axial to radial motion) and the radii of the inlet and outlet. As the flow enters the
compressor, its relative velocity increases so that the relative kinetic energy accounts for
30% - 40% of the work input. Once the flow discharges from the impeller, it enters a
diffuser which decreases its velocity and increases its static pressure. The air is then
delivered to the engine cylinders.
The compressor and turbine for the turbocharger were instrumented in order to
obtain the pressure/temperature rise through the compressor and the pressure/temperature
drop through the turbine. Stainless steel tees were used throughout the turbocharger so
that both temperature and pressure readings could be obtained from only one drilled hole
at each location (Figure 41). A compression fitting was used to join the stainless tubing to
the tee at the pressure branch of the tee. Tygon tubing was used to extend the pressure
transmitter‟s barbed input to the open end of the stainless tubing.
68
Figure 41: Thermocouple and Pressure Transmitter Sketch
69
For the compressor inlet, a bulkhead fitting was used to attach the stainless steel
tee to the sheet metal pipe (Figure 42):
Figure 42: Compressor Inlet Attachment
70
For the compressor outlet, the intake manifold was drilled and tapped then the
stainless steel tee was attached with a nipple (Figure 43):
Figure 43: Compressor Outlet Attachment
71
For the turbine inlet, the exhaust manifold was tapped into and a nipple was used
to attach the stainless steel tee (Figure 44):
Figure 44: Turbine Inlet Attachment
72
For the turbine outlet, a half-coupling was welded to the sheet metal pipe and a
nipple was used to join the coupling and stainless steel tee (Figure 45):
Figure 45: Turbine Outlet Attachment
73
The stainless tubing is essential for the turbine side of the turbocharger since the
exhaust temperatures can raise up to and above 1,000 °F. The stainless tubing allowed for
free cooling of the pressurized air inside the tubing so that the pressure transmitter did not
burn up. A simple fin calculation was completed to determine the minimum length,
diameter and thickness of stainless tubing to use for the turbine side of the turbocharger;
the results are shown in Figure 46:
800
Temperature Distribution of a Long Fin of
Stainless Steel Tubing
700
d=1/8 in, w=0.02 in
Temperature ( C)
600
d=1/8 in, w=0.049 in
500
D=1/4 in, w=0.01 in
400
D=1/4 in, w=0.083 in
300
200
100
0
0
0.1
0.2
0.3
0.4
0.5
0.6
Distance (m)
Figure 46: Stainless Steel Fin Calculation
74
0.7
0.8
0.9
Type K thermocouples and Dwyer differential pressure transmitters (Figure 47)
were used throughout the turbocharger.
Figure 47: Dwyer Differential Pressure Transmitters
75
A final photograph of the entire completed turbocharger instrumentation is shown
in Figure 48.
Figure 48: Entire Turbocharger Instrumentation
5.1.3 Steam System
A Vaporphase ECXSH-3080-2.0 steam generator was selected to produce steam
up to 92 psia for use in the steam absorption chiller during the summer and a hot water
converter during the winter. The steam generator is rated at 20 kW (67,575 Btu/hr), and
will generate 30.4 kg/hr (67.1 lbs/hr) at 6 bar (87 psig). The specifications for the unit are
shown in Table 7.
76
Table 7: Vaporphase Steam Generator Specifications
In this setup, a small 16 kW (4.5 refrigeration tons) absorption chiller was
installed and integrated with the chilled water system. Absorption chillers are key
components in a CCHP system to cool space in buildings. They generate chilled water to
provide a portion of a building's space cooling needs. Conventional refrigeration chiller
units use electricity to drive the water cooling with compressors. Absorption chillers are
instead driven by thermal energy. The combination of heat recovery equipment and heat
driven absorption chillers provides significantly increased overall energy efficiency. The
chiller is integrated with the chilled water system to flow through a system of mullions
77
surrounding the office area. The selected chiller runs on the water-LiBr (lithium
bromide) refrigeration cycle, and was provided by the Broad Air Conditioning Company.
The steam generator is shown in (Figure 49), the engine exhaust gas can either
flow into the steam generator or bypass it and exit straight out of the stack. If entered into
the steam generator, the exhaust gas flows through tubes transferring heat to the water
enclosed by the steam generator (shown in Figure 50). The exhaust gas flows through the
length of the steam generator and then makes a 180 degree turn and flows through again
(through another set of tubes), thus it is called a two-pass steam generator.
Figure 49: Steam Generator
78
Figure 50: Inside View of Steam Generator
5.1.4 Domestic Hot Water
The PHE-Type P4-17-TL plate and frame heat exchanger was made by ITT Heat
Transfer. It operates in parallel with the engine radiator to transfer the thermal energy
from the engine coolant to the ventilation system for regenerating the solid desiccant or to
the IWESS for space heating. The heat exchanger specifications are shown in Table 8,
and an engineering drawing of the heat exchanger is located in the appendix (Figure 65).
79
Table 8: Heat Exchanger Specifications
The coolant heat exchanger is capable of recovering 64,000 Btu/hr at a flowrate of
28 gallons per minute. A photograph of the heat exchanger before the insulation was
installed is shown in Figure 51.
80
Figure 51: Coolant Heat Exchanger Before Insulated
5.2 Procedure and DAQ
For each fuel, we ran tests at 6 kWe, 12 kWe, 18 kWe and 25 kWe. The recovered
heat and the electrical energy created by the engine allows us to measure the overall
efficiency of the system and the pressure, temperature measurements across the
turbocharger allow us to calculate the isentropic efficiency of the compressor.
Before turning the engine on, we input a set point for the steam pressure in the
steam generator (87 psig during the summer and 30 psig during the winter). If the water
level sensor notices that the water level is low inside the steam generator then the feed
water valve is opened.
The transfer switch monitors the three-phase voltage, phase angle and frequency.
Once a power set point is selected, the engine controller adjusts the load on the engine.
81
When starting the engine, the engine controller varies the speed of the engine (1800 +/- 2
RPM) to force the generator to be in-phase with the grid. Once the grid and generator are
in-phase the engine ramps up to the set power production level. If there are any problems
with the generator, the transfer switch will open the generator breaker to protect the grid.
Once the engine is started, the coolant pump automatically pumps as it is belt
driven. A controller starts in heat recovery mode allowing 100% of the coolant to go
through the heat exchanger, and closes the valve to the remote radiator (located outside of
the building). The radiator allows us to dump heat to the outside air if we do not need to
recover heat with the heat exchanger. Once the coolant outlet temperature reaches 180 °F,
the water pump starts and circulates water through the heat exchanger. If the coolant
entering the engine is greater than 165 °F then the radiator valve is opened allowing some
of the heat to exit through the radiator.
All of the measurements are fed into a central data acquisition system. This
system was developed and maintained by Automated Logic. This system can be accessed
and controlled remotely; Figure 52 is a printout showing a list of all the systems. Reports
can be made directly within the software and the type of data for each report is input
manually then the report can be exported to excel for analysis. Figure 53 shows the
system interface.
82
Figure 52: CCHP Monitored Systems
83
Figure 53: Automated Logic Interface
5.3 Experimental Results
5.3.1 Engine
For a baseline to compare with biodiesel, a number of tests were run with the No.
2 diesel. The engine performed closely to the manufacturer‟s specifications (shown in
Table 9) with some issues related to the turbocharger.
84
Table 9: Measured Engine Data vs. Manufacturer Specifications
System
Air System
Max. Temp
Rise, Amb. to
Inlet
Engine Air
Flow (32 kWe)
Intake Manifold
Pressure
Fuel System
Fuel
Consumption (6
kWe)
Fuel
Consumption
(12 kWe)
Fuel
Consumption
(18 kWe)
Fuel
Consumption
(25 kWe)
Fuel
Consumption
(32 kWe)
Cooling
System
Engine Heat
Rejection
Coolant Flow
Exhaust
Exhaust
Temperature
Max. Allowable
Back Pressure
Specification
Data
Notes
15 °F (8 °C)
92 °F (33.3 °C)
Measured after
Compressor (at 25
kWe)
99 CFM (2.8
m^3/min)
83 CFM
Measured at 25 kWe
9 psig (64 kPa)
2.3 psig
Measured after
Compressor (at 25
kWe)
4.7 lb/hr (2.1 kg/hr)
2.1 kg/hr
Verified with Weigh
Tank
7.0 lb/hr (3.2 kg/hr)
3.6 kg/hr
Verified with Weigh
Tank
9.8 lb/hr (4.4 kg/hr)
4.8 kg/hr
Verified with Weigh
Tank
13.3 lb/hr (6.1 kg/hr)
6.4 kg/hr
Verified with Weigh
Tank
17.9 lb/hr (8.1 kg/hr)
N/A
Soft load controller's
maximum allowable
power is 25 kWe
1,303 Btu/min (23
kW)
18 kW
Measured at 25 kWe
24 GPM (91 L/min)
5 GPM
Spec. assumes radiator
is attached to engine
963 °F (517 °C)
1,000 °F (538
°C)
Measured at 25 kWe
14 in-H2O
Measured between
steam generator and
engine exhaust at 25
kWe
30 in-H2O (7.5 kPa)
85
The engine specifications were based on the engine running at 32 kWe, however
the engine cannot exceed 25 kWe due to the programming of the soft load controller. If a
set point for a higher power output is selected, then the controller generates a fault and
shuts down the engine.
Also, notice that the maximum temperature rise from ambient to intake manifold
is 15 °F, and the measured temperature rise is 92 °F. This is due to the low pressure boost
from the compressor (2.3 psig when the engine is rated at 9 psig). A large portion of the
work transferred to the compressor from the turbine is being absorbed in the form of heat
and not compressing the air. The engine exhaust temperature is much higher in our case
(1,000 °F at 25 kWe versus 963 °F at 32 kWe) proving that the turbine is not using the
exhaust gases from the cylinders effectively. Given that the fuel flows and air flow match
closely to the engine specifications, the turbocharger efficiency needs to be assessed in
order to find out if it is properly sized for this engine.
5.3.2 Turbocharger
The Mollier, or enthalpy-entropy diagram provides a convenient way to represent
the ideal and real compressor and turbine processes. The isentropic portions of the
process are represented as vertical lines, and the real processes include internal losses (or
gains of entropy). The ideal and real processes act between the same two pressures; the
compressor and turbine Mollier diagrams are shown in (Figure 54).
86
Figure 54: Mollier Diagrams for Compressor and Turbine
The isentropic efficiency provides the means to calculate the work transfer of an
ideal machine. For an isentropic compressor acting on an ideal gas, the relative pressure
(Pr) can be related between the two states as:
The relative pressure is a function of temperature only, therefore the relative pressure can
be found for the first state. Using the relation above, the relative pressure can be
calculated for the second isentropic state. Now, the enthalpies for all three states can be
87
found, leading to the calculation of the isentropic efficiency of the compressor. The
isentropic efficiency of the compressor is defined as the ratio of the work input required
to isentropically raise the pressure of a gas to a specified value to the actual work input,
shown in the following equation:
This equation can be simplified to:
The turbine isentropic efficiency is defined as the ratio of the actual work output of the
turbine to the work output that would be achieved if the process were isentropic, or:
This equation can also be simplified to the following:
Table 10 shows the average temperatures, pressures and air flows found when
running the engine with No. 2 diesel fuel.
88
Table 10: Average Turbocharger Results for No. 2 Diesel
Power Comp_in Comp_out Comp_ΔP Turb_in Turb_out Turb_ΔP Air Flow
(kWe) T (°F)
T (°F)
(psi)
T (°F)
T (°F)
(psi)
(kg/s)
6
12
18
25
89.67
91.68
94.41
94.38
115.55
131.12
156.00
186.38
0.39
0.91
1.56
2.32
518.18
712.96
931.54
1,121.39
463.14
633.45
835.66
998.63
1.73
2.14
2.74
3.97
0.0361
0.0373
0.0398
0.0433
Table 11 shows the average temperatures, pressures and air flows found when
running the engine with the soy biodiesel.
Table 11: Average Turbocharger Results for Soy Biodiesel
Power Comp_in Comp_out Comp_ΔP Turb_in Turb_out Turb_ΔP Air Flow
(kWe) T (°F)
T (°F)
(psi)
T (°F)
T (°F)
(psi)
(kg/s)
6
12
18
25
85.46
87.96
91.64
91.38
110.21
126.48
149.77
175.45
1.71
2.15
2.84
3.67
478.96
675.63
892.91
1,085.40
430.89
598.66
791.63
968.03
0.78
1.73
2.99
4.80
0.0370
0.0377
0.0399
0.0428
Overall, from the No. 2 diesel and soy biodiesel data it seems that the
turbocharger is more effective when using the soy biodiesel. The turbine has a higher
change in pressure across it when running with the soy biodiesel, and that absorbed work
is generating a higher change in pressure for the compressor. Thus, more air is being
packed into the engine cylinders. The next graph (Figure 55) shows the compressor
isentropic efficiencies for the No.2 diesel and soy biodiesel fuels.
89
Figure 55: Isentropic Compressor Efficiency
The isentropic compressor efficiency decreases with increasing power for the
biodiesel and increases with increasing load (until 18 kWe) for the No. 2 diesel fuel.
90
In Figure 56, the turbine isentropic efficiencies are shown. The specific heat ratio,
k, was constant, at 1.4, for the compressor efficiency calculations and fluctuated between
1.389 and 1.352 for the turbine efficiency calculations.
Figure 56: Isentropic Turbine Efficiency
91
The isentropic turbine efficiency performs the same as the compressor efficiency
with the soy biodiesel (decreasing with increasing power) but the isentropic turbine
efficiency steadily drops with increasing power when the engine was running with No. 2
diesel fuel.
Overall the turbocharger performance is not efficient possibly because the device
is not matched well to the needs. In order to diagnose any problems with the
turbocharger, it is necessary to analyze the compressor map. For the compressor to
operate efficiently, it must operate at a high angular speed. The compressor map shows
lines of constant compressor efficiency, constant corrected mass flow rates (x-axis) and
pressure ratios (y-axis). For a given pressure ratio, a specific mass flow rate is required
for the compressor to operate smoothly and effectively.
92
Figure 57 shows the compressor map for the turbocharger installed in this engine:
Figure 57: Compressor Map
The mass flow rates with pressure ratios from our data (Figure 58) show that this
engine is operating outside of the specified range given by the compressor map. The
minimum pressure ratio for the compressor is about 1.39 (while the highest pressure ratio
in the data shows 1.25). The mass flow rates are also very low, therefore the isentropic
efficiencies above reflect the fact that the turbine is not rotating fast enough to provide
enough boost for the given mass flow rate.
93
Pressure Ratio vs. Mass Flow Rate
1.3
Pressure Ratio (P/Patm)
1.25
Soy
#…
1.2
1.15
1.1
1.05
1
0.035
0.036
0.037
0.038
0.039
0.040
0.041
0.042
0.043
0.044
Mass Flow Rate (kg/sec)
Figure 58: Pressure Ratio vs. Mass Flow Rate
5.3.3 Steam System
The heat recovered by the steam generator is calculated based on the first law of
thermodynamics:
Figure 59 and Figure 60 show the heat recovered by the steam generator for No. 2 diesel,
and soy biodiesel, respectively.
94
Exhaust Heat Recovered by Steam
Generator
25
Heat Transfer (kW)
20
15
10
Nov. 2007
Dec. 2007
5
Mar. 2008
0
0
5
10
15
20
25
Time from Engine Start (hr)
Figure 59: Heat Transfer to Steam Generator (No. 2 Diesel Fuel)
The No. 2 diesel fuel data was taken during the months of November, December
and March. In November and December, the engine was run at 25 kWe and took about 10
hours to reach steady state. The heat recovered from the engine exhaust about 16.5 kWt
in November and 16 kWt in December. In March, the engine was operated at all settings
(25 kWe, 18 kWe, 12 kWe, and 6 kWe). The amount of heat recovered was 17.3 kWt,
11.9 kWt, 7.5 kWt, and 4.0 kWt, respectively.
95
Exhaust Heat Recovered by Steam
Generator
25
Heat Transfer (kW)
20
15
10
Oct. 9, 2008
Oct. 15, 2008
5
Oct. 22, 2008
0
0
5
10
15
20
25
Time from Engine Start (hr)
Figure 60: Heat Transfer to Steam Generator (Soy Biodiesel)
The soy biodiesel data was taken during the month of October. The engine was
run at 25 kWe, 18 kWe, 12 kWe, and 6 kWe. The data for the 6 kWe setting will not be
analyzed because an accurate fuel flow reading was not possible at that setting. At 25
kWe, the average heat recovered was about 17.5 kWt, at 18 kWe the average heat
recovered was 12.5 kWt, and at 12 kWe the average heat recovered was 10.4 kWt.
5.3.4 Domestic Hot Water
The heat recovered by the No. 2 diesel fuel and the soy biodiesel during the tests
are shown in Figure 61 and Figure 62, respectively.
96
Heat Transfer to Water in Coolant Heat
Exchanger
25
Heat Transfer (kW)
20
15
Nov. 2007
10
Dec. 2007
5
Mar. 2008
0
0
-5
5
10
15
20
25
Time from Engine Start (hr)
Figure 61: Heat Transfer to Water Loop (No. 2 Diesel Fuel)
With the No. 2 diesel fuel, the heat recovered from the coolant during November
and December (at 25 kWe) was about 18 kWt. In March, the heat recovered was 18.6
kWt at 25 kWe, 14.4 kWt at 18 kWe, and 11 kWt at 12 kWe. The data for 6 kWe is
ignored since very little heat transfer occurred between the coolant and water loop.
97
Heat Transfer to Water in Coolant Heat
Exchanger
30
25
Oct. 9, 2008
Heat Transfer (kW)
20
Oct. 15, 2008
Oct. 22, 2008
15
10
5
0
0
5
-5
10
15
20
25
Time from Engine Start (hr)
Figure 62: Heat Transfer to Water Loop (Soy Biodiesel)
The soy biodiesel data shows that at 25 kWe, about 17.7 kWt is recovered, at 18
kWe 13.8 kWt is recovered, and at 12 kWe 10 kWt is recovered. Again, the 6 kWe data is
ignored since there is very little heat transfer between the water and coolant. Also, notice
that during the October 15th data set, the heat transfer drops to zero for a long period of
time (about 20 hours). During the time period, there was not a significant demand for hot
water so the coolant was routed to the radiator so that the heat could be dumped outside
to the atmosphere.
98
CHAPTER 6
RESULTS AND DISCUSSION
In the present work the engine performance characteristics with a straight
vegetable oil (SVO) and two types of biodiesels were compared with that of No. 2 diesel
fuel. The SVO used was rapeseed (Canola oil) and the two types of biodiesels used were
produced from two different feedstocks, Soybean oil and waste vegetable oil (WVO)
from restaurants. Each of the fuels was tested in a 6 HP single-cylinder compression
ignition (CI) engine. Additionally, Soy biodiesel and No. 2 diesel fuel performances were
compared using a 42 HP compression ignition engine in a combined cooling and heating
(CCHP) facility at Carnegie-Mellon University. The results of this study can be
summarized as follows:
There was no significant variation in BHP with the use of SVO, Soy biodiesel or
No. 2 diesel fuel for all full throttle conditions used here. However, the BHP
began to drop for the WVO as the engine load reached 30 psig BMEP and
continued to drop to about 5 percent less than the other fuels at full load (64 psig
BMEP). This is due to the lower heating value of the WVO resulting in less work
produced per cycle which was verified by plotting P-ν diagrams for all fuels.
The mechanical efficiencies (BHP/IHP) were similar with all fuels, but the Soy
biodiesel stayed consistently above the other fuels across all engine loads. The
mechanical efficiency was lowest when fueled with the SVO. At full load, all
fuels converged to 52 percent mechanical efficiency.
99
As expected, the Brake Specific Fuel Consumption (BSFC) was best across all
loads with No. 2 diesel fuel due to its higher heating value. The SVO BSFC was
17 percent higher, the Soy biodiesel 14 percent higher, and for the WVO 5
percent higher at full load when compared to No. 2 diesel fuel. The BSFC was
lower with the WVO than with the Soy biodiesel even though the WVO has a
lower heating value than the Soy biodiesel. The BSFC for SVO was low at low
load and high at high load.
The Soy biodiesel had the worst thermal efficiency across all loads out of all the
fuels. It was 7 percent less than No. 2 diesel fuel at full load. The thermal
efficiency was 4 percent less with SVO compared to No. 2 diesel fuel at full load
and the thermal efficiency increased 8 percent with WVO when compared to No.
2 diesel fuel. The heating value with of the WVO is 14 percent less than the
heating value of No. 2 diesel fuel, accounting for the increase in thermal
efficiency. The better efficiency may be due to higher amount of Oxygen and
better combustion.
The SVO had the highest volumetric efficiency across all loads; this is associated
with its high air flow into the engine cylinder. The other three fuels converge to
about 76 percent volumetric efficiency at full load.
The RPM severely drops off for the WVO as the engine load increases because
the engine load (torque) is proportional to the horsepower divided by the RPM.
Since the WVO produces less horsepower at higher loads (25 psig BMEP and
100
higher), the RPM of the engine must drop to generate the same torque. At full
load, the RPM decreased 5 percent compared to No. 2 diesel fuel.
The net exhaust temperature is slightly higher with the Soy biodiesel when
compared to the No. 2 diesel fuel and SVO across all loads. At low load the WVO
exhaust temperature was similar to the other fuels but at higher load (when the
BHP was less than with the other fuels) the exhaust temperature decreased.
The waste heat was not recovered in the single-cylinder diesel engine but based
on potential recovered heat from the exhaust and horsepower, the CHP efficiency
is lower with the Soy biodiesel across all loads and consistently higher with No. 2
diesel. For the WVO, the efficiency reached extremely high values at high load.
The problems with the SVO at low load (low fuel flow, low exhaust temperature
and increased air flow) were due to the fact that the engine was at a lower temperature
therefore the viscosity of the SVO was high (75 cSt at room temperature, 20°C). At rated
conditions (3600 RPM, 5.8 HP), the fuel temperature at the fuel injector pump inlet is
40°C and lower at low load. The engine manual states that at low temperatures if running
with a high viscosity fuel, it may result in fuel delivery problems, injection pump
seizures, and poor injection nozzle spray atomization. Poor spray atomization results in a
drop in combustion efficiency. This is shown in the P-ν diagrams for SVO at low load
where during the blow down stroke, the pressure builds inside the cylinder. Therefore, the
flow of the exhaust gases out of the engine cylinder is inhibited.
101
Future work for the 6 HP single-cylinder CI engine should include tests under part
throttle conditions and some upgrades in instrumentation are needed in order to optimize
results. The upgrades needed are:
The dynamometer water valve needs to be retrofitted so that engine load can be
controlled more precisely and in a repeatable manner for various fuels.
The throttle knob needs better calibration (a set of notches along the housing) so
that wide open throttle as well as part throttle conditions can be easily
demarcated.
A fuel flow meter should be installed on the fuel line so that an accurate
measurement can be made. This will be better than using the weight method to
determine fuel flow rate.
An electric fuel pump can be installed on the engine to ensure that there is a
constant supply of fuel for the fuel injection pump.
Fuel analysis should be done for every sample used according to ASTM
standards.
An estimate of the pressure drop associated with the intake system should be
calculated.
The ambient pressure and humidity need to be measured to correct for the inlet air
conditions.
102
The Soy biodiesel was also compared with No. 2 diesel fuel in a combined
cooling heating and power (CCHP) environment at CMU. The experimental results are
shown in Table 12 and Table 13.
Table 12: CCHP Results with No. 2 Diesel Fuel
Fuel
Plant Power Exhaust Coolant
CCHP
Test Date Input Power Output Recovery Recovery Efficiency
(kWc) (kWe) (kWe)
(kWt)
(kWt)
(%)
11/02/07
12/07/07
03/05/08
80.73
80.73
80.73
60.55
45.41
26.49
4
4
4
4
4
4
25
25
25
18
12
6
16.47
16.14
17.32
11.93
7.48
4.08
17.96
17.47
18.64
14.39
11.18
-
68.66%
67.64%
70.56%
66.60%
58.69%
22.94%
Table 13: CCHP Results with Soy Biodiesel
Fuel
Plant Power Exhaust Coolant
CCHP
Test Date Input Power Output Recovery Recovery Efficiency
(kWc) (kWe) (kWe)
(kWt)
(kWt)
(%)
10/09/08
10/15/08
10/22/08
66.21
46.67
85.30
66.21
85.30
66.21
46.67
4
4
4
4
4
4
4
18
12
25
18
25
18
12
13.87
9.60
16.80
12.00
19.28
14.41
10.39
13.92
10.91
18.21
13.77
17.29
13.63
10.35
63.13%
61.08%
65.66%
60.06%
67.49%
63.50%
61.59%
The Soy biodiesel performed well in the CCHP system. At full load (25 kWe) the
fuel energy input to the engine increased by 5 percent but the heat recovered by the steam
generator increased by 8 percent. The heat recovered from the coolant heat exchanger
decreased by 2 percent. The total CCHP efficiency at 25 kWe decreased by 3 percent
when using Soy biodiesel instead of No. 2 diesel.
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At 18 kWe, the fuel energy input to the engine increased by 9 percent when using
Soy biodiesel instead of No. 2 diesel fuel but the heat recovered by the steam generator
increased by 13 percent. The heat recovered from the coolant heat exchanger decreased
by 4 percent. Therefore, the CCHP efficiency at 18 kWe decreased by 7 percent when
compared using Soy biodiesel instead of No. 2 diesel.
At 12 kWe, when switching from No. 2 diesel fuel to Soy biodiesel, the fuel
energy input to the engine increased by 3 percent but the heat recovered by the steam
generator increased by 34 percent. The heat recovered from the coolant heat exchanger
decreased by 4 percent. The total CCHP efficiency at 18 kWe increased by 4.5 percent
when compared using Soy biodiesel instead of No. 2 diesel.
From these results, it is clear that No. 2 diesel fuel is the better option for a CCHP
system instead of Soy biodiesel. The extra heat recovered from the steam generator for
the Soy biodiesel (from the higher fuel flow rate) does not compensate for the extra fuel
energy input to the system and loss of heat recovered from the coolant heat exchanger.
Soy biodiesel did perform well enough to be considered as a possible fuel for a CCHP
system and may still be preferred since it is renewable.
The turbocharger for the 42 HP four-cylinder CI engine is oversized. The boost
pressure across the compressor is extremely small due to the turbine rotating at a slow
RPM. The exhaust gas flow rate is not high enough to power the turbine and the resulting
operating position on the compressor map is far from acceptable. If this turbocharger was
properly sized for this engine, the pressure exiting the compressor would be between 24
and 43 psia, and the highest pressure obtained at the compressor outlet for this setup was
18.4 psia.
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Even though the turbocharger is undersized, there is evidence to suggest that the
turbocharger performance increases when using Soy biodiesel versus No. 2 diesel. Table
14 and Table 15 show the results for No. 2 diesel fuel and Soy biodiesel, respectively.
Table 14: Turbocharger Results for No. 2 Diesel
Power Comp_in Comp_out Comp_ΔP Turb_in Turb_out Turb_ΔP Air Flow
(kWe) T (°F)
T (°F)
(psi)
T (°F)
T (°F)
(psi)
(kg/s)
6
12
18
25
89.67
91.68
94.41
94.38
115.55
131.12
156.00
186.38
0.39
0.91
1.56
2.32
518.18
712.96
931.54
1,121.39
463.14
633.45
835.66
998.63
1.73
2.14
2.74
3.97
0.0361
0.0373
0.0398
0.0433
Table 15: Turbocharger Results for Soy Biodiesel
Power Comp_in Comp_out Comp_ΔP Turb_in Turb_out Turb_ΔP Air Flow
(kWe) T (°F)
T (°F)
(psi)
T (°F)
T (°F)
(psi)
(kg/s)
6
12
18
25
85.46
87.96
91.64
91.38
110.21
126.48
149.77
175.45
1.71
2.15
2.84
3.67
478.96
675.63
892.91
1,085.40
430.89
598.66
791.63
968.03
0.78
1.73
2.99
4.80
0.0370
0.0377
0.0399
0.0428
The boost pressure when the engine was run on Soy biodiesel at 25 kWe was 58
percent larger than with No. 2 diesel fuel. At 18 kWe it was 82 percent larger, at 12 kWe
it was 136 percent larger and at 6 kWe it was 338 percent larger. These numbers reflect
the higher compressor efficiency at low load for the Soy biodiesel (70 percent at 6 kWe,
56 percent at 12 kWe, 49 percent at 18 kWe, and 43 percent at 25 kWe). The compressor
efficiency stays around 22 percent when the engine is running on No. 2 diesel fuel
(except at 6 kWe where it drops to 13 percent). The turbocharger is working more
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efficiently with the Soy biodiesel since the exhaust mass flow is higher (due to more fuel
being used).
In conclusion, biofuels can be used in current compression ignition engines quite
effectively. Some biodiesel fuels may even be more effective under certain diesel
applications. Engine performance characteristics with biodiesels are similar to those with
fossil diesel which makes biodiesel fuels an alternative to help overcome the current
energy and environmental crises.
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APPENDIX
Figure 63: Air-Flow Nozzle Chart
110
Figure 64: Recycled Grease Process Heat Exchanger
111
Figure 65: Heat Exchanger Engineering Drawing
112
VITA
Tim Guider received his Bachelor of Science degree in Physics from Moravian
College in May 2007. He started graduate school at Lehigh University in August 2007
and is planning to obtain his Master of Science in mechanical engineering in December
2008.
While conducting research in the area of biodiesel, Tim has also been working for
the Department of Energy‟s Industrial Assessment Center (IAC) located at Lehigh
University. The IAC performs energy assessments for industrial plants in the
Pennsylvania and New Jersey areas, and prepares detailed reports based on
thermodynamic principles to quantify energy savings. The average annual energy savings
is over $150,000 per assessment or about 10-15% of a company‟s total energy bill. He
has been working for the IAC since August 2007 and has participated in well over 20
energy assessments.
For the summer of 2008, he interned at the Electrotechnology Applications Center
(ETAC) located at Northampton Community College in Bethlehem, PA. He performed 8
energy assessments for industrial and commercial clients. He took a lead role in
completing the on-site assessments and contributed greatly to preparing each report. He
also conducted research for various companies using infrared and microwave
technologies, including one project which resulted in the implementation of infrared
heaters to the preheat process of EPDM rubber.
Tim is a member of the American Society of Mechanical Engineers (ASME) and
the American Institute of Physics.
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