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                             -2-

                      Table of Contents
I.   Introduction

II.  Summary
                                                     Page

                                                       3

                                                       5
III. Conclusions
                                                       8
IV.
A.  Volumetric Efficiency Losses
B.  Prevention of Volumetric Efficiency Losses

C.  Optimum Fuel Delivery

Discussion
 8
 8

 8

 9
     A.  Adiabatic Temperature Drop vs.
         Vapor Pressure
     B.  Heat Transfer
References

Appendix I

Appendix II


Appendix III


Appendix IV    -


Appendix V

Appendix VI


Appendix VII


Appendix VIII
                Partial Pressure Calculation

                Temperature Drop With
                Vaporization

                Volumetric Efficiency Change
                Due to Adiabatic Temperature Drop

                Observed Volumetric Efficiency
                Loss

                Heat Transfer Calculations

                Changes in Pumping Work due to
                Temperature Changes

                EPA Memo, Results of Fumigation
                Testing

                Test Data
11
16

22

23


25


35


38

39


44


48

63

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                             -3-

         Methanol Vaporization:  Effects  on Volumetric
          Efficiency and  on  Determination of  Optimum
                     Fuel Delivery System
I.   Introduction

The  Office   of   Mobile  Sources  within   the   Environmental
Protection  Agency  has   studied   and  evaluated  alternative
transportation  fuels  since  its   formation  in  1970.   EPA's
responsibilities   under   the   Clean   Air  Act  also   have
necessitated  a   significant   regulatory   role   dealing  with
transportation  fuels.   In particular,  Section 211  of  the
Clean  Air  Act  requires  EPA  to play  a  key  role  in  the
introduction  of  new fuels  and  fuel  additives.   Perhaps  most
visible  was  EPA's  role   in  the  introduction  of  unleaded
gasoline  to  permit the  use of catalytic  converters on  1975
and  later  model  year  automobiles.    More recently EPA  has
responded to  a growing  interest in the use of  oxygenates  (in
particular  methanol)  for  use  in  motor  vehicles  and  for
blending with gasoline.

As  part  of  this  response to  the  interest  in methanol,  a
program  was  undertaken  to grade  several  fuel  utilization
concepts.  One of  these  concepts  that has been  evaluated  has
been   loosely  labeled   fumigation.   Normally,    the   term
fumigation is used  in  the context of diesel engines when  all
or  a majority  of  the   fuel  is  introduced  into  the  intake
track,  as opposed to injecting  the fuel into the cylinder  (or
pre-chamber).   Generally,   fumigated  diesel   engines   are
unthrottled engines  and rely on  the  fuel delivery  system to
regulate  the  load.   In  our  testing  with  a  throttled  spark
ignition  (SI)  engine  converted  to  methanol  use,  only  a
portion  (20-33%)  of the  fuel was  injected  or  "fumigated"  into
the  intake  track.  Two  fumigation  locations  (see  figure  1)
were used  -  one  significantly  upstream  from  the  throttle
valve,   and  the  other  slightly  downstream of   the  throttle.
Both locations were upstream from the  normal port  injectors
which injected the majority of  the fuel  (for more  details  see
reference 1).

Testing  has  shown  this  upstream  fumigation of methanol  to
cause  a  decrease  in  volumetric  efficiency(1)*.   Volumetric
efficiency  is  usually defined  as a  measure  of the mass  of
air/fuel charge which is  successfully taken into the cylinder
on each  intake stroke.   It  is an  "efficiency" with  respect to
utilization of the  cylinder space,  but  only indirectly  with


*Numbersin parentheses designate References at  end  of  paper.
Reference (1) is  reprinted in Appendix VII.

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                         -4-
     AIR FILTER
  POPT INJECTORS
         \
                THROTTLE
                 VALVE
FUMIGATION
"INJECTORS
(LOGAT ION 2)
              Figure 1

            FUMIGATION
          TEST SET-UP

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                             -5-

respect to  energy  or fuel utilization.  A  loss  in volumetric
efficiency  would  have  as  its  most  direct  consequence  a
reduction  in  power  for  a  given throttle position  and engine
speed,  including  a  reduction  in  maximum  power.   Pumping
losses  would  increase,  causing  a  small reduction  in overall
engine efficiency.   The observed loss in volumetric efficiency
was opposite  of the commonly held  assumption  that vaporizing
the   fuel   would   cool   the  intake  air   and   increase  the
volumetric  efficiency.   Thus,  a small  mathematical  exercise
was undertaken  to  attempt to quantify  the  physical processes
causing the real  effects  to be opposite  of  the  anticipated
ones.  Although  the  results of  this exercise  did  not uncover
new  phenomena  governing  engine operation  (as  a  review  of
Reference  2 will  indicate),  it  did reinforce the  importance
of  existing  but   sometimes  over-looked phenomena  on  engine
operation.   Consideration  of  these phenomena  may  be  more
important for a methanol engine  than for a gasoline engine.

II. Summary

EPA tests  utilizing  fumigation of  methanol  caused  a decrease
in  volumetric   efficiency.    The   apparent  cause   of  this
reversal  in  the  actual   versus   the   expected  result  was
primarily due to heat  transfer effects from the walls  of the
intake  passage.   A  secondary  issue deals  with  the necessity
for the combination  of  the local partial pressure  of  the fuel
and the stagnation temperature  to be on the vapor  side of the
vapor  pressure  curve  for  the  fuel in  question in  order  to
allow  vaporization  to  take place.   The  partial pressure  is
the key  to this secondary effect  and  is  influenced  directly
by  the local  equivalence  ratio.*   In  the cases  discussed,
this  secondary   effect  had  only a minor   influence  on  the
results,  primarily  because  the effective   equivalence  ratio
was   very   lean  at   the  point  of   fumigation  (i.e.,   the
fumigation  injectors received only part of the total  engine
fuel  flow  resulting  in a  much  lower partial pressure  of the
fuel  in the local  area).  If, however,  the  local  equivalence
ratio   at   the   fumigation   injector   had   approached   a
stoichiometric  mixture  or   richer,   the   partial  pressure/
temperature effect would prevent a  significant portion of the
fuel  from  vaporizing.   In  this case  (rich condition),  the
intake  air  temperature depression  due  to  vaporization  would
be limited  to the  level allowed by the vapor pressure  curve
of the substance.

As stated,  our experiments  were  not substantially  effected by
the secondary effect because our local air/fuel  ratios  were
*  Note:    the   Greek  word   "phi"   is  used   to  designate
equivalence ratio throughout the report.

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                             6

very lean.  Under  these  conditions,  all or nearly all  of  the
fuel was evaporated.  On the other hand,  even  though  the lean
mixture  allowed   vaporization,   the   temperature   depression
available  from  this vaporization was  limited  because  of  the
small mass  of  fuel involved.  Under  these  conditions,  normal
component  temperatures  of  the  intake air  duct-work  appeared
to  be  sufficient  to  cause  the   initially  cooled  mixture  of
fuel and air to heat back  up to  the  original  temperature (due
to  the  temperature differential  between  the wall  and  the
newly  cooled  mixture).   Since  the   air  returned   to  its
original  specific  volume  (at  its original temperature),  and
the vaporized fuel had a much  larger  specific  volume  than  the
liquid  fuel,  the  volume  of   the  mixture  (air  plus  gaseous
fuel)  was  greater  than the   volume  of  intake  air  with  no
fuel.  Because  the volume  flow rate  (Q)  of the engine would
remain  the same  for  the  same  manifold  vacuum,*  an  intake
charge of  this gaseous  fuel-air mixture would contain less
air than an intake charge of air without the fuel.

This decrease in air  flow  when using a  fuel-air mixture will
manifest itself in  a  low volumetric efficiency  (VE).   In  our
case, if heat transfer is  sufficient  to reheat the  mixture to
its  original  temperature,  then  the   theoretical   loss   in
volumetric  efficiency  (for phi =  0.8)  is  around  2.3  percent
for  location  1 and around  3.8  percent  for  location  2 (see
Table 1).   The  average  loss  obtained from  an  analysis  of  the
test results  is 2.8 percent  for location  1  and 4.1  percent
for location 2.
                           Table 1

                    Volumetric Efficiency

                              Location 1           Location 2

Calculated loss
(Including Heat Transfer)        2.3%                 3.8%

Measured loss (average)          2.8%                 4.1%
*For a  complete derivation,  see  Appendix III.  It  works  out
that the  total  mass flow and  total gaseous volume  flow vary
only slightly with  changes  in  density  for a constant pressure
drop.  What changes is the  portion  of  the total  that is fuel,
the  remainder   is  the  change  in  inlet  air  flow.   This  can
readily  be seen  by  comparing  the component volumes  of  a
stoichiometric  mixture per  pound  of fuel  in Tables  AIID-1 to
3 in Appendix II (air component =Mvl, fuel component = Mxv2).

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                             -7-

Recognizing  that  the  average  test  results  mask  some  data
scatter, we are  left with  the  fact  that  fumigating  a fraction
of   the   total  fuel  flow   caused  a  loss   in  volumetric
efficiency.   The  theory  suggests  that  heat  transfer  which
reheats  the  charge  to  the  original  temperature and  expands
the  fuel  volume can explain  a majority, if  not all,  of  the
difference  between  the   expected  results   and  the  actual
results.  Any  differences  not explained by  heat transfer  are
probably due  to experimental error, or  some other  phenomonea
not yet fully understood.

In  principle,  it   should  be  possible  to  reduce  the  heat
transfer down  stream of  the  fumigation point.   This  might be
accomplished,  for  example, by  insulating the inner  surface of
the  intake  track.   However,   the  temperature  drop would  be
limited  by  the dew point  of  the  mixture  strength.   For
example a carbureted  system  at stoichiometric  conditions  and
15 in. Hg MAP  would require  a mixture temperature  above 50F
to  maintain   100%  vapor.   (Note:  At this  temperature,  the
effect of  fuel  expansion  would  cause a  theoretical  loss  in
volumetric efficiency of slightly  greater  than 6%.)   On  the
other hand,  if one were  willing  to tolerate a  portion of  the
fuel  in  liquid  form,  there  could  be  an   equilibrium  point
where the mix  temperature  is depressed  sufficiently  to allow
an    improvement   in    volumetric    efficiency.     However,
distribution of  the liquid  fuel  could become a problem.   If
partial  fumigation with  an  insulated  manifold  was   used  in
combination with port  injection  for cylinder  distribution,  a
small  increase  in volumetric efficiency  might be  possible
(our  testing  suggests  2  to  4  percent - see table  3).   This
increase  might  only be   available  under  certain  operating
conditions due to  changes  in the  equilibrium condition of  the
mix during different operating regimes.   Therefore,  while  it
might  be  possible  to  obtain   some  benefit  with  partial
fumigation,   it would require  careful design  to insure  that
such  potential  improvements  would   occur   at the   design
operating point, since it  seems  possible that the  off-design
points would incur losses in volumetric efficiency.

The   results   of  this   experimental  work   and  theoretical
analysis allowed us to make  the  following  generalizations  and
conclusions.

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                             -8-

II. Conclusions

    A.   Volumetric Efficiency Losses

A loss in volumetric efficiency when mixing  methanol  with air
in  the  inlet  system  (compared to  air only)  is primarily  a
result of heat transfer effects warming the mixture.

    B.   Prevention of Volumetric Efficiency Losses

By  awareness  and  proper  manipulation of  the  following  five
factors  a  potential  loss  in  volumetric  efficiency  can  most
likely be avoided.

    1.   The partial pressure  of  the fuel is governed  by the
    mixture strength  (f/a),  the  local manifold  air  pressure
    at the  time  of vaporization,  and the portion of  the  fuel
    permitted  to   be   vaporized   by  the   partial  pressure
    characteristics at the local temperature.

    2.   The  temperature  of  the  mixture  after  the  heat  of
    vaporization  is included  is  relative  to  the  degree  of
    vaporization,  the  partial  pressure of  the  fuel, and the
    amount  of  the  fuel.   The  temperature of  the  mixture  is
    also affected  to  some  degree  by  the initial  stagnation
    air temperature.

    3.   The heat  transfer to  the mixture is governed  by the
    difference between  the wall  temperature and the mixture
    temperature,   but   is  also  influenced by   the  amount  of
    liquid fuel coming in contact with the wall.   The effects
    of the liquid  fuel contact may be substantial.

    4.   The length of the intake  passage  from  the  point  of
    fuel  introduction  to  the  cylinder  over which  the  heat
    transfer can take place is of concern.

    5.   The time  during  which heat  transfer  can  take place
    also affects the  results.   The  time can be  influenced  by
    engine speed,  load  (i.e.,  throttle position),  geometry  of
    the  intake  tract  (i.e.,  velocity  in   the  system),  or,
    distance (i.e., length of tract).

    C.   Optimum Fuel Delivery

From these  factors, it  appears that the best location  to add
fuel  when   considering   volumetric   efficiency   (VE)   with
methanol is directly  into  the  cylinder when the  intake valve
is open.  Direct injection would probably be  the  best(2), but
port injection near the  intake valve would be  expected to  be
a  better   choice   on  a  cost/benefit   basis   than  direct

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                             -9-

injection.  Port  injection,  of course,  would be  expected  to
be much better than carburetion for  VE.   Because  of potential
side   benefits   of   improved   driveability   (less   manifold
wetting,  etc.)  and  possibly  better  emission  control,  port
injection  should  also  be  better   than  carburetion  on   a
cost/benefit  basis.   If  carburetion  is  used,  however,  more
effort  should  probably  be  directed  at  developing  a  more
uniform  wet-flow  cylinder  distribution  for   the  methanol
intake  manifold  than has  been found  to be  necessary  for  a
gasoline  manifold.   This analysis  of  methanol  evaporation
versus   vapor  pressure  and  temperature   suggests   that
sufficient  manifold  temperature  will not  be available  (and
may never be  available)  under  enough  conditions  to  avoid wet
flow  mal-distribution  problems   (e.g.  cold  start,  warm-up,
etc)  if  a  normal  gasoline manifold/carburetor combination  is
used when using methanol.

Considering  the  qualitative  aspects  of this  analysis,  one
might predict that phased port injection  would perform better
than normal port injection.  Such  a  prediction would be based
on  the  assumption that  a  majority  of the fuel  would  be
expected  to be evaporated  within  the  cylinder   when  phased
port  injection  was used.   Such evaporation  when  the  intake
valve  is  open  would  be  expected  to  improve  volumetric
efficiency because the  greater volume to surface  area within
the cylinder  (compared  to  the intake tract) would tend  to
limit the heat transfer to the evaporated mixture.  Secondly,
any evaporation after the intake valve is closed  would reduce
the work of compression by cooling the mixture.   Finally, any
heat  transfer  that did  occur  would  cool  the cylinder  wall,
which would  tend  to  reduce  the  load  on the cooling  system.
These assumed physical  reactions  could allow the potentially
cheaper  phased  port   injection   system  to duplicate  the
performance  of  a  more  expensive  direct  cylinder  injection
system.

IV. Discussion

The original focus of the fumigation testing  was  to ascertain
any potential  benefits  that  might  be available  due to the
high  latent  heat   of  methanol.  An  analysis  in   Appendix  VI
suggests that if we were to  cool  dry  air  (i.e. no fuel) down
to  the  lowest  fuel  dew-point  temperature  experienced  in
testing,  we  could reduce  the  pumping  work  by  roughly  2
percent  for  a constant  volume flow  rate  or in  the case  of
constant mass flow rate, 15 percent.   We  hypothesized  that  by
vaporizing only a  small  fraction  of  the  fuel upstream  of the
port injectors, the high  latent  heat  in  methanol  would  allow
some improvement in  volumetric efficiency,  and possibly some
reduction in the pumping work.  The  test  results  contradicted
this hypothesis.

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                             -10-

In  order  to  investigate  this  problem/   the  basic  precepts
behind   volumetric   efficiency  were   examined.    Volumetric
efficiency  can  be  defined  as  the  amount  of  intake  air
ingested  (at  STP)  expressed  as a  percentage  of  the  swept
volume  of  the  engine.   Several  factors can  influence  the
volumetric  efficiency at  wide  open  throttle;  examples  are
camshaft  design,   valve   size,  exhaust  restricition,  intake
path  size  and  length,  and resonant  effects  due  to  intake
geometry.   At part  throttle  all  of  the above factors  are
modified  by the throttle  valve.   In  our  experiments all  of
the basic factors were  held constant  since  the same  engine
was used.   Also, all  tests  were  conducted  at  the same RPM and
the same low level  of torque  (which  required part  throttle
operation).   The  comparisons  of  volumetric  efficiency  were
all  made  at  the  same  manifold  vacuum  level  (i.e.,   same
throttle position).

Under  these conditions,   only  two basic  factors are readily
apparent   that  could cause  the   volumetric   efficency  to
decrease  when fumigation was  added.  One is  that the  basic
air ingestion characteristics of the engine were changed with
the addition  of  the  fumigated  fuel.  The other is  that the
fumigated  fuel displaced  some  of the  air  that  would  normally
be  ingested if  the fuel  were  not  there.   In the  first  case,
it  is hard  to imagine  how  adding  fuel  to  the  air stream  could
change the  basic  air ingestion characteristics  of  the  engine
under  the moderately  throttled  condition as  occured in our
tests.    At   wide  open   throttle,   factors   such   as   poor
cylinder-to-cylinder   distribution,   or   a   change  in   the
resonant  effects due  to  the additional fuel mass might  affect
the  volumetric  efficiency,  but  the  power  would  also  be
expected  to vary, hence  dissolving the comparison at  constant
power.   The  second   consideration,  the  one  of  fuel  volume
replacing air volume  is  more palatable,  but  only  if  the fuel
is vaporized.  In liquid  form, the volume  displacement of the
fuel  is   so small  that  the effect  on volumetric  efficeincy
would be expected to be pratically negligable.

Therefore,  if  only the  vaporized  fuel has  the  potential  to
cause  the  majority  of  observed  effects,  then  the  factors
influencing vaporization must be  evaluated  to  determine  if
these factors could have been  present  in  sufficient magnitude
to  create  the  observed   effects.    The  factors   affecting
vaporization are  temperature and pressure.  As will  be  shown
in  the  following  sections,  these  factors  have a  dramatic
influence  on  the  amount of  vaporization that   takes place.
The first step taken to evaluate the potential volume  increase
of  the  fuel was  to identify  the expected temperature of the
intake air  after  a  given quantity of  methanol  was  evaporated
(regardless  of  other influencing   factors).   The  following
sections  address  the  determination of this temperature  drop,

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                             -11-

and how  the  vapor  pressure of the fuel,  in  concert  with heat
transfer effects, influence the  final fuel volume.

    A.   Adiabatic Temperature Drop vs. Vapor Pressure

If heat  input  is  assumed to  be equal  to  zero  for  the time
being, equation AII-1  (ref.  2)  in  Appendix  II  provides  an
estimation of  the temperature  drop  of  the  intake  air.   The
temperature drop predicted by  the equation  is  governed by the
portion of the  fuel  evaporated  and the  local  fuel-air ratio.
Since the first evaluation looked at  temperature drop with no
heat  transfer,  the  term Q  was  set  at  zero  for  the  first
analysis.

 (AII-1)     AT = [(x) (F) (HLG)  + (Q)] /  (1-F + xF)  (Cp)

In  order  to  determine  the  amount  of  fuel  that  can  be
evaporated  under  any  given   condition,  the  vapor  pressure
versus   temperature   characteristics   of  the   compound   in
question  must  be  known.   Data  on  these characteristics  was
obtained  from  reference 3 for methanol  (see Table  2),  and a
curve of  that  data  was  constructed (see  Figure  2).   The next
step  in  order  to  use  this  data  would  be  to  determine  the
partial pressure  of  the  methanol  under  the  local conditions
which  are controlled  by  the local  fuel-air  ratio  and  the
portion of the  fuel  evaporated  (e.g.,  25%,  50%,  75%, 100%). A
range  of  partial   pressures  can  be   calculated  for  each
fuel-air   ratio (see  Appendix I  Section IIIC).   Using  these
partial  pressures,  and  the  curve in  Figure  2,  a  dew  point
temperature  can  be  estimated  such   that   if  the' mixture
temperature  were  below  this  temperature  (and  at   the  same

                           Table 2

             Vapor  Pressure Curve for  Methanol(3)

                     mm HgA         C

                        1         -44.0
                        5         -25.3
                       10         -16.2
                       20         - 6.0
                       40         4-5.0
                       60         +12.1
                      100         +21.2
                      200         +34.8
                      400         +49.9
                      760         +64.7
*F = f/a, HLG = heat of vaporization, x = portion evaporated

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                                            -12-
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                             -14-

partial  pressure),  it  would be  expected  that  any  methanol
vapor  would  condense  back  into  liquid droplets  until  new
equilibrium  conditions  were  satisfied.   The  same  assumed
evaporative  percentages  used  to   calculate   the   range  of
partial pressures can also be substituted into  the  intake air
temperature drop  equation  (e.g.,  AII-1), to obtain estimated
adiabetic  intake  air  temperatures  (Tl-  ATI  in Appendix  II
Section IIB).

If  the  estimated  intake air  temperature from  equation  AII-1
is  warmer  than  the  dew  point  temperatures  (i.e.  Tl-ATl
greater   than  TDEW,    Appendix   II,   Section   IIB)   then
vaporization could be assumed to take place.

In our  case,  where  we  introduced only 20 percent of  the fuel
at  location 1  (Figure   1)  upstream  of  the normal  injection
system  (the remainder  from  the port  injectors),  all of  the
fuel  can  be  evaporated at  this upstream location  due  to
partial  pressure  effects  (Location  1,  Appendix II,  Section
IIB).   However,  the  richer  of  the  two   equivalence  ratios
indicates   that  total   vaporization   is   marginal   without
additional heat since the depressed  intake  air  temperature is
effectively  the same as the dew point  temperature.   If  we
look  at  Location  2  (Figure 1)  where  33% of the  fuel  is
introduced  upstream  of  the  port  injectors    (Appendix  II,
Section IIB), the  leaner of the  two equivalence ratios  would
have  marginal  evaporation,   while   the  richer  ratio  cannot
vaporize  all of the  fuel from partial pressure  effects  alone
(only about 78% will evaporate).  The visualization of this
effect of  incomplete evaporation  can be seen by plotting the
calculated  intake temperature versus the vapor  pressure  as in
Figure 3.

The result  of this dew-point  limiting  temperature  can also be
seen  in  the volumetric  efficiency changes  (CVE)  in Table  3
(positive  is gain,  negative  is  loss).   At  location 2, we see
a  4.1 percent  improvement  in  volumetric  efficiency with  75
percent  of  the  fuel  evaporated.   Whereas  at  100  percent
evaporated,  the volumetric  efficiency  improvement drops  to
only 2.4 percent.   This  is because the temperature  of the mix
must be  raised  from the 75  percent  evaporation dew  point  to
the dew point temperature necessary  to achieve  evaporation of
100% of the fuel.

Comparing  the   calculated  volumetric efficiency  improvements
in Table  3  to the  actual results in Table  4, we see  there is
a  sizeable  difference.   At an  equivalence  ratio of  0.8,  the
calculated  results for location 1 predict a 2.5  percent  gain,
while  the  actual  results  show  a  2.8  percent  loss.    The
numbers are a   3.8 percent  gain,  and a  4.1 percent  loss  for
location  2.  Obviously the  results do not  match,  and further
investigation was  in order.

-------
                             -15-
Phi
, **
1.0
0.8
                           Table 3

            Volumetric  Efficiency Change  "Without"
           Heat Transfer from the Engine Structure
          X
         .25
         .50
         .75
        1.00

         .25
         .50
         .75
        1.00
CVE* (%)-Location 1

        0.7
        1.5
        2.4
        2.9

        0.6
        1.2
        1.8
        2.5
CVE* (%)-Location 2

        1.2
        2.6
        4.1
        2.4

        0.9
        2.0
        3.1
        3.8
*CVE is calculated by Equation AIII - 22 in Appendix III
**Represents overall equivalence  ratio,  the local equivalence
ratio at location 1 was  20%  of  th overall  ratio,  and 33.3% of
the overall ratio at location 2.
                           Table 4

                Average* Volumetric  Efficiency
                	  Change  (Test  Results)	
     Phi
     i **
     0.8
CVE (%)-Location 1
                       -2.8
CVE (%)-Location
                              -4.1
*See Appendix IV.

**Several   different   equivalence    ratios
Generally they ranged between 0.7 and 0.9.
                                             were
                                                     tested,

-------
                             -16-

    B.   Heat Transfer

By noting  the  probability that only  marginal  evaporation was
likely to  take  place  under some conditions  (from  the  20% and
33%  fuel  split at  the  fumigation  injectors),   the  physics
seemed to  imply that as more  fuel  was added to the fumigation
injectors, the  amount of  the  fuel  that could be vaporized was
reduced.   The   apparent   reason   for  this  was   that   the
additional fuel flow  (either  from a  larger  flow split  or  a
richer  fuel  air ratio)  caused  the  partial  pressure of  the
fuel  to   increase  which  in  turn   raised   the  dew  point
temperature  necessary  for  evaporation.  in  the  extreme,  if
all of the fuel were to be introduced at the front end of the
intake manifold  (as in  a   carburetor)  the partial pressure of
the fuel would  be considerably greater (from 2.8 to 4.5 times
greater*)  than  the  cases  analyzed.   In  order  for  total
evaporation to  occur under these  conditions,  the final intake
air temperature  cquld  not be  below 45 to  50F (assuming 15
in. HgA  MAP).   However,  since vaporizing the  total  fuel flow
would depress  the  inlet air  temperature approximately 300F,
and assuming an initial inlet  air  temperature  of around 80F,
a discrepancy  of  some  265F  exists  that must  be made  up by
heat  transfer   from  some  source   before the  fuel  can  fully
vaporize.  Equation AII-1  indicated that the addition of heat
during   evaporation   would   cause   the  final   intake   air
temperature depression to  be  less  than in  the  absence of heat
addition.    Therefore,   heat   transfer  becomes   a  logical
parameter  to  be   investigated in resolving  the  difference
between the expected results and the measured results.

Since heat  transfer  would  raise  the  temperature  of  the  mix,
it would be  useful  to  assume  several recovered charge-air
temperatures  in order  to  identify  if further   investigaton
would be fruitful.  Then we  can  perform  the  heat  transfer
calculations  to  identify  if  there   could  be  a  sufficient
temperature  differential   to   create  the   assumed  recovery
temperatures.   The  results in  Table  5  speak  for themselves.
The volumetric  efficiency losses   from  heat  transfer resemble
the actual losses observed  (Table  4)  much  closer  than  the
calculated gains  in  Table  3.   Also,  if we were  to carburet
the  entire  amount  of  fuel,  as   opposed   to  the  fractional
fumigation used in these  tests, a  sizable  loss  in volumetric
efficiency would occur  (assuming   sufficient heat  transfer to
vaporize all of the fuel).
*Compare the mole fraction  (mf  100)  for  4  injectors (4 total)
which would  be all  of the  fuel  to  the  mole fraction  for  1
injector  (5  total)  and 2  injectors  (6  total)  for  the  two
equivalence ratios in Appendix I Section C.

-------
                             -17-
It  is  interesting  to  note that  the  calculated  percent  loss
for volumetric efficiency  remains  constant  for apparently any
temperature above the  original  air  temperature (assumed to be
80F for location 1, and 90F  for  location  2).  The mechanics
of   this   anomoly    (constant  VE   loss   with   increasing
temperature)  are  that  the  comparisons  are  made  with  all
components at  the  same recovery temperature.   Therefore,  the

                           Table 5

             Voluemtric Efficiency Change "with"
           Heat Transfer from the Engine Structure
        Resulting  in an Assumed Charge-Air Temperature
Assumed Charge-
Air Temperature

-Partial Fumigation*
     80F
    100F
    120F

     80F
    100F
    120F

-Carbureted*

    100F
  i *
Phi
 1.0
 1.0
 1.0

 0.8
 0.8
 0.8
 1.0
 0.8
 CVE (%)
Location 1
  -2.9
  -2.9
  -2.9

  -2.3
  -2.3
  -2.3
  12.8
  10.5
 CVE (%)
Location 2
  -3.9
  -4.7
  -4.7

  -3.0
  -3.8
  -3.8
  13.7
  11.4
*Theequivalence  ratio  for partial fumigation  represents the
overall  equivalence ratio,  the  local  equivalence  ratio  at
location  1  is  20%  of  the overall  ratio,  and  33.3%  of  the
overall  ratio  for location 2.   For the  carbureted  condition
the local  and the  overall equivalence  ratios  are  the  same.
For  both   the   fuel  delivery  senerios,  the   pressure  at
location 2 is approximately 15 in. Hg of manifold vacuum.
total  volume  of  fuel  and  air  at  120 F is  compared to  the
volume  of  air  with  no  fuel  at  120F,   not  at  80F  (the
original  inlet  temperature).   The comparison at  the  elevated
temperatures  is  made  because  it  is  assumed  that  any  heat
transfer  that  would  heat  the  mixture to some  temperature
above the original air temperature  would also heat  the intake
air  in the baseline  case  (which did  not  have  fuel in  the
mixture)  to  the same temperature.   The reason the base  case
temperature would not be higher,  is  that in the base  case the
temperature gradient between the  duct wall  and  the  intake air

-------
                             -18-

is  very  small.   Hence  little  heat  is  transfered.   However,
when methanol  is  injected,  the intake air  is  cooled creating
a  reasonably  large  temperature  gradient  which  allows  much
more rapid  heat transfer.   Once  the cooled fuel-air  mixture
has returned to the  base case, then any temperature gradient
that would  have affected  the base case  would  also  affect the
fuel-air mixture  in a  similar  fashion  because  the specific
heats of air and  methanol vapor are  very similar.   Hence, the
conclusion  that the  volumetric efficiency  change between air
alone  and   the  fuel-air mixture  must be based  on the  same
recovery temperature.

Because  the  investigation   of heat  transfer  effects  began
after   the   tests   were   run,   many   of   the   temperature
measurements  that  would be  useful  for   the  heat  transfer
calculations were not  available.  Therefore,  many of  these
temperatures were  assumed based on other factors.   The  issue
here  is  to identify if  rough  heat transfer  approximations
with reasonable temperature assumptions can account  for any
of the discrepancy between  the expected  volumetric  efficiency
change and  the  measured change.  As  it  turned  out,  even with
extremely   unsophisticated   heat   transfer   approaches,   heat
transfer  does   account  for  a  majority  if not  all  of  the
difference between the expected and the observed.

The  first  assumption was  that  the  only heat transfer  that
occurred,  occurred  between the  intake  tract wall and  the
vapor phase methanol-air mix.  By using this  assumption,  we
were able to apply equation (13.2 - 25)  from  reference  4 for
convection  in  tubular  heaters  (this  equation   is repeated  as
AV-1.  The  various boundary  assumptions  used for  the analysis
are listed  in Appendix VI.

The first case  analyzed was for  the single injector upstream
of  the   inlet  to  the  EFI  system (ie.  upstream of the  air
filter).   The   total distance  from  single injector  to  the
throttle  body  mounted  on  the   intake air   collector  was
approximately 80 inches.

Since  nearly   all  of  this   distance   was   surrounded  by
atmospheric  temperature,  it  was assumed  that at   least  the
initial  portion of the  inlet  system was at  room temperature
(approximately  80F).   It  was   further  assumed   that  the
temperature gradually increased along  the  length of the duct
to  a  final  temperature  of   160F.   This temperature  rise  is
considered  to  be an upper  limit for estimation  purposes for
the following  reasons.   At  about the 40 percent  point  in the
intake tract,  the  intake  air passes  through a large box-like
structure that houses  an  air filter.   In  leaving the  air
filter box,  the intake  air  passes through a spring  loaded air
meter device.   The box  and  air  meter are metal and  are in the

-------
                             -19-

vicinity  of  the exhaust manifold  (within 10-12  inches).   It
is  assumed  some radiative  heat transfer  occurrs  to  the  air
filter/metering  system.   However,   no  attempt  was  made  to
correct  for  any of  the  potential heat  transfer  improvements
due  to   the   proximity  of  the  exhaust   manifold   to  the
filter/metering box, to  the  increased surface  area  in the  air
filter,  or  to  the  increased  residence  time  in  the  air  box
other than to  assume the above  temperature gradient along  the
duct.  Following  the air  box  was a  long section  of stamped
steel duct work in  close proximity to the cylinder  head.  And
finally,   a  short   section  in  front  of  and  including  the
throttle  valve was  heated  by engine  coolant.   The  coolant  was
maintained  at  about  180F  on  the  engine  dynamometer  test
stand.

By  applying  these  boundary  conditions  to the  tubular heater
equation, the  length of  the  intake tract  necessary  to recover
the   original   air  temperature    (after   the   temperature
depression due  to  evaporation)  varies between  a  low estimate
of 85 inches  (phi = 0.8) to  a high value  of  137 inches (phi =
1.0).  The variation in  these  calculated estimates  is due to
a  range   of  assumed hydraulic  diameters  of  the  irregularly
shaped intake  tract.   The measured  distance on  the  hardware
was approximately 80 inches.

The  discrepancy between the measured  tract  length  and  the
calculated  range   can   be   attributed   to  several  reasons.
First, mentioned  before,  is that  more  heat   transfer  could
have occurred  in the air filter box  than  was accounted for in
the calculations.   Second, and  most  likely,  is the  estimation
procedure itself.   All of  the  temperatures were estimated  and
not  measured   (note:  if 100F had  been  used for  the  final
wall  temperature  (To2)  instead  of  160F,  the  calculation
would  predict  slightly less   than  a  doubling  in  length).
Probably  more  important,   however,   is  the  procedure  only
considered   convective    heat   transfer   to   the   gaseous
methanol-air  mixture.   Heat  transfer with  a change  in  phase
(fuel evaporation)  can  be very  complex.   The  effects  on  the
system due  to  convective  heat  transfer  on  any liquid  fuel
droplets  that  might have impinged on the wall  could  have  had
a  noticeable   influence  on  the  total  heat   transfer  rates.
This  effect,   however,   was  not  considered   in   the  rough
approximation.  Finally, there  is the  additional  possibility
that some other unknown parameter  influenced  the  volumetric
efficiency results.

Even  though  some  people   might  consider   the  discrepancy
between  the  calculated and measured  results  to  be  minor
(i.e.,  substantially less  than one magnitude  of  difference),
investigation  into  the   results  at  Location  2 might support
one  of   the  suggested  reasons  for  the  discrepancy    more

-------
                             -20-

specifically,  the  issue of  neglecting  the influence  of  heat
transfer  to liquid  fuel  droplets.   At Location  1,   a  small
amount  of  fuel  was  sprayed   into   the   center  of  a  large
diameter  duct  (4.5  inches)  approximately  6  to  8  inches
upstream from  the entrance to  the  normal  inlet duct-work.   At
Location  2, 65  percent more  fuel  was sprayed  into an  air
collector about 2  to 3 inches from  the entrance  of  four  1.35
inch  intake runners.   Because of  the  physical  set-up,  more
liquid  droplets would  be  expected  to impinge  on   the  duct
walls at Location 2  than at  Location 1.  If the heat transfer
effects from the droplets were significant,   one would expect
the  simple   calculation  procedure  (i.e.,   neglecting droplet
effects)  to  show  less  agreement  at  Location  2  than  at
Location 1.

The simple  calculation  predicts  a path length  for  Location 2
of  between 31.9   and  32.2  inches   (phi   =   0.8   and   1.0
respectively).    The   approximate   measured  distance  is  14
inches.  The over  estimation by the  simple approach  is  about
128%  in this  case   (note:   the over  estimation value  would
increase if lower  wall temperatures  were  used).   At Location
1, the  over  estimation  for the smallest assumed duct diameter
is only about  6%,  whereas the length of  the  largest assumed
diameter  is   over   estimated  by   around  58%.    From   this
comparison,   it is  difficult  to  scientifcally  say   that  the
simple  approach  is  more  accurate   at Location  1  than  at
Location 2.  Hence,  the hypothesis  that  liquid fuel droplets
impinging on the intake walls  increased the heat transfer can
not be  completely confirmed.   However, since  the  convective
heat  transfer   coefficient   (h)   for   boiling   liquids   is
approximately  100  times  that  for  forced  convection  of gases,
and about   5 times  that  of  gases  for  viscous fluids(4)(5),
there certainly  is  room for  engineering  judgement  to expect
that  any fuel droplets  on   the  walls  would  substantially
enhance  the heat   transfer.   Enhanced heat   transfer  would
shorten the  effective  length of the  intake track necessary to
recover the original  intake  air  temperature,  thus  bringing
the calculated lengths more in line with the actual hardware.

In conclusion,  the effects of  heat  transfer appear  to play an
important  part  in  explaining  the  difference  between  the
expected VE results and  the  actual  results.   One may  argue
that  the  estimated  component  temperatures are too  high,  or
that  the   study  should  have  investigated  the  effects  of
droplets  on the  walls  more  thoroughly,   but  in  the  final
analysis,  the  aggregate evidence  says  heat transfer  plays  a
role in the  volumetric efficiency  of the  engine when methanol
is introduced  into  the inlet  track.  Furthermore, because of

-------
                             -21-

the  difference   in   the  heat   of   vaporization*,   the  heat
transfer  role  appears to be  more dramatic  for  methanol than
gasoline,  and  therefore  this  role  should  be  given  more
consideration in  the design of methanol fuel delivery systems.
*Obert(6) provides a  procedure  to calculate the  dew  point of
gasoline  (a  multi-component  mixture).   Between  10%  and  90%
evaporated, the dew point  temperature  is  very  similar between
gasoline  and  methanol.  However,  due  to  differences  in  the
heat  of  vaporization  and  the  specific heats, methanol  will
have  an  adiabatic  temperature  drop  eight  times  (on  the
average) greater than  for  gasoline.  Therefore,  to maintain a
similar dew point mixture  temperature  in  the intake manifold,
the heat  transfer  required  for  the methanol  engine  would be
substantially greater than for the gasoline engine.

-------
                             -22-
                          References

(1)  Results of  Methanol Fumigation  Investigation,  EPA memo,
April  22,   1983,  B.  Michael  and  W.  Clemmens   to  C.  Gray,
reprinted in Appendix VII.

(2)  The Internal  Combustion Engine  in Theory  and  Practice,
Volume  1,  second edition,  C.F.  Taylor,  The  MIT Press,  MIT
Cambridge,  Massachusetts, pages 183-185.
(3)  Chemical  Engineers  Handbook,
Perry, 1973, p. 3-56.
5th  edition,   Robert  H.
(4)  Transport Phenomena,  R.  Byran Bird, Warrren  E. Stewart,
Edwin  N.  Lightfoot,  John Wiley  and  Sons,  1960,   Library of
Congress (60-11717),  Figure 13-2,  p  400,  and pp 398-407 which
is referenced  to  E.N. Sieder  and  G. E.  Tate,  Ind.  Eng. Chem.
"28", 1429-1435 (1936).

(5)  Principles  of  Heat  Transfer,   second  addition,  Frank
Kreith,  International Textbook  Company,  Scranton,  PA, 1967,
p. 15.

(6)  Internal Combustion  Engines  and Air Pollution, Edward F.
Obert,  Intext  Educational  Publishers,  New  York,   1973,  pp.
254-263.

(7)  Aeronautical   Vest-Pocket   Handbook,   Pratt   &  Whitney
Aircraft, Eleventh Edition, Eighteenth Printing, May 1966.

-------
                                     -23-

                                  Appendix I

                         Partial Pressure Calculation
I.  Basic Equation
    Phi = 1
(AI-1)    CHsOH + 1.5(02 + 3.73 N2 + .04 Ar)  -* C02 + 2H20 + 1.5 (3.73N2 + .04Ar)

II. Molecular Weights

         CH30H = 32.043           N2 = 28.014

             02= 32               Ar = 39.948

III. Partial pressure (Pv) of Methanol

     A.   Equation for mole fraction with 100% vaporization


(AI-2)      (mflOO)  =  (z)+(1.5)+(i.g)(1.73)+(1.5)(.04)


     where z = % of the total fuel used at specific location

               times 10~2

          mf = mole fraction

     or

                           (z)
(AI-3)  (mflOO)  =
                    (z)  + (7.155/PHI)
    where           Phi = (f/a)  actual/(f/a)  stoichiometric = equivalence ratio

    B.    Equation for mole fraction wih x% vaporized


(AI-4)  mfx =   (Z)(x) + (7.155/PHI)


    where   x      = % of z amount of fuel that is vaporized times 10~2

                     (e.g. 25,  50, 75,  100)

-------
                                 -24-
     C.   Values from equation AI.4.



         Configuration        Z      Phi    mf 25    mf  50    mf  75    mf  100
Cl

C2

C3
*

. 1
1
4
4
. 2
2
4
4
. 4
4
injector*
injector*
injectors"1"
injectors*
injectors*
injectors*
injectors"1"
injectors"1"
injectors"*"
injectors"1"
(5
(5
(5
(5
(6
(6
(6
(6
(4
(4
total)
total)
total)
total)
total)
total)
total)
total)
total)
total)
.2
.2
.8
.8
.333
.333
.667
.667
1.0
1.0
1
.8
1
.8
1
.8
1
.8
1
.8
.007
.006
.027
.022
.012
.009
.023
.018
.034
.027
.014
.011
.053
.043
.023
.018
.045
.036
.065
.053
.021
.015
.077
.063
.034
.027
.065
.053
.095
.077
.027
.022
.101
.082
.044
.036
.085
.069
.123
.101
Fumigation Injector
Port Injector
D.
Partial
Pressure
(pv)





(AI-5)         (pv)  = (mfx)(local  absolute  pressure)

-------
                             -25-

                         Appendix II

             Temperature  Drop with  Vaporization

I.   Injector Location

    A.   Location 1

         1.   Upstream single injector with 4 port injectors.

         2.   Fuel flow split:  20%  at Location 1, 80%  at  the
              normal port injectors

         3.   PI  = 29  in.  HgA  (736.6  mm  HgA) ,  Tl  =  80eF
              (26.7C)

         4.   Overall phi ratios examined = 1.0,  0.8

    B.   Location 2

         1.   Two  air   collector   injectors   with  4   port
              injectors.

         2.   Fuel flow split:  33.3% at Location  2,  66.7%  at
              the normal  port injectors

         3.   P2 = 15 in.  HgA(381.0 mm HgA),  T2 = 90F
              (32.2C)

         4.   Overall phi ratios examined = 1.0,  0.8

II.  Calculations

    A.   Temperature drop (AT)  Equation for x% evaporated(2)

(AII-1)   AT = C(x)(F)(HLG) + (Q)]/(1-F + xF)(Cp)

    where

         x  = % evaporated = 0.25, 0.5, 0.75, 1.0

         F  = (f/a) = (flow split)(.155)(phi)

         Cp =  .240 for F  less  than 0.5; = .245  for  F greater
         than 0.5 (assumption for approximate calculations)

         HLG = heat of vaporization = 474 BTU/lb.
          AT = F
         Q   = zero for the first stage of analysis

-------
                             -26-
    1.   Location 1
phi
1.0
1.0
1.0
1.0
.8
.8
.8
.8
2.
phi
1.0
1.0
1.0
1.0
1.0
.8
.8
.8
.8
X
.25
.5
.75
1.0
.25
.5
.75
1.0
Location 2
X
.25
.5
.75
.78
1.0
.25
.5
.75
1.0
AT1(C)
8.7
17.3
25.7
34.0
6.9
13.8
20.5
27.2

AT2(C)
14.7
29.1
43.0
44.7
56.6
11.7
23.1
34.3
45.3
   f/a = (.2)(.155)(phi)

Tl-ATl(C)      Pv(mm)    TDEW(C)*
18.0
9.4
1.0
-7.3
19.8
12.9
6.2
- .5
f/a = ( .333)
T2-AT2(C)
17.5
3.1
-10.8
-12.5
-24.4
20.5
9.1
-2.1
-13.1
5.2
10.3
15.5
19.9
4.4
8.1
11.8
16.2
(.155) (phi!
Pv(mm)
4.6
8.8
13.0
13.3
16.8
3.4
6.9
10.3
13.7
-25.6
-15.9
- 9.8
- 5.9
-27.8
-19.4
-13.9
- 9.1

TDEW(C)*
-27.2
-18.2
-12.5
-12.1
- 8.6
-31.2
-21.7
-15.9
-11.7
* Calculated from:  Log10 Pv = [-0.2185(A)/K]+B
                           A = 8978.8   Range:  -44C to 224C
                           B = 8.639821
                           K = K

   Source:   CRC  Handbook  of  Chemistry  and  Physics,  53rd
edition, 1972-1973.

    B.   Specific Volume of Air

    1.   Location 1

              v = RT/p

              where

                   p = (29.0) in hg (70.73) = 2051.2 psf

                   T = (459.7 + 80) - ATI = (539.7 - ATl)R

                   R = 53.345 ft-lb /R

                   v = specific volume = cu ft/lb air

-------
                        -27-




  phi          x           T (air, R)    v(cu ft/lb air)
^"-_ --
1.0
1.0
1.0
1.0
1.0

1.0
1.0
1.0
.8
.8
.8
.8
.8

.8
.8
.8
0
25
50
75
100
with heat transfer
100
100
100
0
25
50
75
100
with heat transfer
100
100
100
539.7
524.0
508.6
493.5
481.1
to 80,100,
539.7
559.7
579.7
539.7
527.3
514.9
502.9
490.8
to 80, 100
539.7
559.7
579.7
14.04
13.63
13.23
12.83
12.51
and 120F
14.04
14.56
15.08
14.04
13.71
13.39
13.08
12.76
,and 120F
14.04
14.56
15.08
2.   Location 2
          v = RT/p




          where




               p = 15 in HgA (70.73) = 1061.0 psf




               T = (459.7 + 90)-AT2 = (549.7  - AT2)R




               R = 53.345 ft-lb/R

-------
                    -28-

phi           x          Tair
1.0
1.0
1.0
1.0
1.0
with heat
1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8
with heat
0.8
0.8
0.8
0
25
50
75
100
transfer
100
100
100
0
25
50
75
100
transfer
100
100
100
C. Specific Volume of
549.7
523.2
497.3
472.3
476.2
to 80, 100
539.7
559.7
579.5
549.7
528.6
508.1
487.9
470.6
to 80, 100,
539.7
559.7
579.7
Fuel
27.6
26.3
25.0
23.7
23.9
,and 120F
27.14
28.14
29.15
27.6
26.6
25.5
24.5
23.7
and 120F
27.14
28.14
29.15

1. Location 1
     v  = (RT/p) vapor

where

     p  = (29.0) in hg (70.73) = 2051.2 prf

     T  = (459.7 + 80) -ATI = (539.7 -ATl)R

     m  = 32.043

     mR = 1544 ft-lb/R
     v  = specific volume = cu ft/lb of vapor

-------
                             -29-

1.  Case 1

         phi
1.0
1.0
1.0
1.0
1.0

1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8

0.8
0.8
0.8
Location
0
25
50
75
100
with heat
100
100
100
0
25
50
75
100
with heat
100
100
100
2
539.7
524.0
508.6
493.5
481.1
transfer
539.7
559.7
579.7
539.7
527.3
514.9
502.9
490.8
transfer
539.7
559.7
579.7

                                                 12.68
                                                 13.15
                                                 13.62

                                                 12.68
                                                 12.39
                                                 12.10
                                                 11.81
                                                 11.53
                                                 12.68
                                                 13.15
                                                 13.62
              v = RT/p

         where

              p  = 15 in HgA(70.73) = 1061 psf

              T  = (459.7 + 90)-AT2 =  (549.7 -  AT2)R

              m  = 32.043

              mR = 1544

-------
                         -30-

     phi
1.0
1.0
1.0
1.0
1.0

1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8

0.8
0.8
0.8
Compute
0
25
50
75
100
with heat
100
100
100
0
25
50
75
100
with heat
100
100
100
mass Volume
549.7
523.2
497.3
472.3
476.2
transfer
539.7
559.7
579.7
549.7
528.6
508.1
487.9
470.6
transfer
539.7
539.7
579.7

                                           25.0
                                           23.8
                                           22.6
                                           21.4
                                           21.6
                                           24.51
                                           25.42
                                           26.33

                                           25.0
                                           24.0
                                           23.1
                                           22.2
                                           21.4
                                           24.51
                                           25.42
                                           26.33
D.

     V = (mass air)(vl)+(mass fuel)(x)(v2)+(Mass fuel)(1-x)(v3)

     where

          mass air  = 6.452 Ib (phi = 1)
          mass fuel = Location 1: 20% (1  Ib)  =  .2  Ib for
                                  phi of 1.0;  =.16  Ib for
                                  phi of 0.8

                    = Location 2: 33.3%(1  Ib)   =  .333  Ib
                                  for phi  of  1.0;  =  .266
                                  Ib for phi of 0.8

                    = Carbureted :  1 Ib  for phi = 1.0,
                                   0.8 Ib for  phi =  0.8

     x = % of fuel evaporated = 0,  .25,  .50, .75, and 1.0
     vl= Specific volume air (cu ft/lb)
     V2= Specific volume evaporated fuel (cu ft/lb)
     V3= Specific volume liquid fuel = .0201 cu ft/lb
     V = Cubic feet of mixture per pound of fuel consumed
     vc= Specific Volume of gaseous portion of mixture
     VG=  V/[(mass air) + (x) (mass fuel)]

-------
            -31-

       Table AII-D-1

Specific  Volume of Partially
 Fumigated Mix at Location 1

phi
1.0
1.0
1.0
1.0
1.0

1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8

0.8
0.8
0.8

X
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
0
.25
.50
.75
1.00
with
1.00
1.00
1.00

Mvl
90.59
87.94
85.36
82.78
80.71
heat transfer
90.59
93.94
97.30
90.59
88.46
86.39
84.39
82.33
heat transfer
90.59
93.94
97.30

Mxv2
0
.62
1.20
1.74
2.26
80,
2.54
2.63
2.72
0
.50
.97
1.42
1.84
80,
2.03
2.10
2.18

M(x-l)v3
.004
.003
80.33
.001
0
100, 120F
0
0
0
.0032
.0024
.0016
.0008
0
100, 120F
0
0
0

V
90.59
88.56
96.56
84.52
82.97

93.13
96.57
100.02
90.59
88.95
87.36
85.81
84.17

92.61
96.04
99.48
Specific
Volume of
Mix (vc)
14.04
13.62
13.21
12.80
12.47

14.00
14.52
15.04
14.04
13.06
13.37
13.06
12.73

14.01
14.53
15.05

-------
            -32-

       Table AII-D-2

Specific Volume of Partially
Fumigated Mix at Location 2
phi
1.0
1.0
1.0
1.0
1.0

1.0
1.0
1.0
0.8
0.8
0.8
0.8
0.8

0.8
0.8
0.8
X
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
0
.25
.50
.75
1.00
with
1.00
1.00
1.00
Mvl
178.08
169.69
161.30
152.91
154.20
Mxv2
0
1.98
3.76
5.34
7.19
heat transfer to
175.08
181.56
188.05
178.08
171.62
164.53
158.07
152.91
8.16
8.46
8.77
0
1.60
3.08
4.44
5.69
heat transfer to
175.08
181.56
188.05
6.52
6.76
7.00
M(x-l)v3
.0067
.0050
.0033
.0017
0
80, 100,
0
0
0
.0054
.0040
.0027
.0013
0
80, 100,
0
0
0
V
178.08
171.67
165.06
158.25
161.40
and 120F
183.24
190.03
196.82
178.08
173.22
167.61
162.51
158.60
and 120F
181.60
188.32
195.05
Specific
Volume of
Mix (vc)
27.60
26.27
24.94
23.61
23.79

27.01
28.01
29.01
27.60
26.57
24.45
24.43
23.61

27.03
28.03
29.03

-------
                             -33-

                            Table  AII-D-3

                  Specific Volume  of  Carbureted Mix
                 with Complete Fraction of Fuel and
                  a  Charge-Air  Temperature of  100F


phi      v      Mvl     Mxv2    M(x-l)v3     V        Vc       VB
Location 1
1.0
0.8
1.00
1.00
93.92
93.92
13.15
10.52
0
0
107.06
104.43
14.37
14.40
14.56
ii
Location 2
1.0
0.8
1.00
1.00
181.56
181.56
25.42
20.33
0
0
206.98
201.90
27.78
27.84
27.60


-------
                    -34-

               Table AII-D-4

   Specific Volume  of  Partially Fumigated
Mix (all locations) vs. Fraction Evaporated

Phi
1.0
11
II
II
ii
.8
M
ll
II
tl
with
1.0/80F
1.0/100F
1.0/120F
0.8/80F
0.8/100eF
0.8/120F

X
0
.25
.50
.75
1.00
0
.25
.50
.75
1.00
Location
VB
14.04
M
11
11
II
14.04
II
II
M
II
heat transfer (assumed
100
II
ti
100
II
II
14.04
14.56
15.08
14.04
14.56
15.08
1
Vc
14.04
13.62
13.21
12.80
12.47
14.04
13.70
13.37
13.06
12.73
charge
14.00
14.52
15.04
14.01
14.53
15.05
Location
VB
27.60
M
II
II
II
27.60
ii
M
ii
M
temperature)
27.60
28.14
29.15
27.60
28.14
29.15
2
Vc
27.60
26.27
24.94
23.61
23.79
27.60
26.57
25.45
24.43
23.61

27.01
28.01
29.01
27.03
28.03
29.03

-------
                             -35-

                         Appendix III

             Volumetric Efficiency Change Due to
             _ Adiabatic Temperature Drop

In  order  to compare  the  effects of  the  adiabatically cooled
charge  on  volumetric  efficiency,   it  would  be  useful  to
perform  this  comparison  at  a  constant  manifold  pressure
(which  is  also consistent with  the  analysis  in Reference 1).
If  we consider the Bernoulli equation of the form;
      1)  0.5d1(V1)2 + P-L = 0.5d2(V2)2 + P2*

we can rearrange it to,

(AIII-2)  PI - ?2 = 0.5d2(V2)2 - 0.5d1(V1)2

If  we consider  condition  (1)  to be  at  ambient  conditions,
then  the  term  (P^  -  P2 )  would be  our constant  manifold
vacuum  (MV) ,  and  the  V^   term   would  be  zero.   Therefore
Equation AIII-2 would become

(AIII-3)  MV = 0.5d2(V2)2

Now we can  evaluate  the effects of charge-air  cooling  on the
velocity (V) term by  forming  a  ratio of  the baseline case (B)
to the cooled case (C).
    I-4)  MVC/MVB = [0.5 w(V2)2]c/[0.5 w(V2)2]B

Since MVB = MVc we can change (AIII-4) to,

(AIII-5) C(V2)2]C/C(V2)2]B = W
Also recognizing that the weight  density  (w)  can be expressed
as  the  inverse of  the  specific  volume,  we  have  Equation
AIII-6  which  says  the   velocity  change   in  the  system  is
inversely proportional to the  square  root  of  the ratio of the
specific volumes.

(AIII-6) [V2]c = CV2]B (vc/vB)0-5

The mass  flow  change  due to  the charge  cooling  can  now  be
calculated  from the  information  on  the  velocity  change  and
the specific  volume  change.   This  mass  change  can  then  be
related  to a  change  in volume  flow  at  the   inlet   to  the
induction   system   which  is  at   constant   temperature  and
pressure.  First, the mass rate (M) can be described as,
*d = mass density = weight density/gravitational constant
   = w/g.
 v = velocity.

-------
                             -36-

(AIII-7)  M = (d)(V)(A)

Next,  we  can look  at the  change in  mass  from  the  baseline
     due to charge cooling  (MC) by forming a ratio,

      8)  MC/MB = CdVA]c/[dVA]B

Cancelling  the  areas  (A)  in AIII-8,  and substituting  for YC
from AIII-6, we have
      9)  MC/MB =  (dcVB)(vc/vB)0.5/(dBvB)

Substituting for d (d = w/g) and cancelling VB we have

      10) MC/MB = wc (vc/vB)'5/wB
Substituting the specific volume  (v)  for  the  density (w)  (w =
1/v),

(AIII-11) MC/MB =  (vB/vc)(vc/vB)0-5

Equation  AIII-11  describes   the   change   in  mass  flow  at
constant manifold  vacuum  due  to charge cooling  in terms  that
we  have calculated  in  Section  D  of  Appendix  II  (specific
volume).  Note  that  the  specific volume of the  cooled charge
(vc)  is the  specific  volume  for  a  mix  of vaporized  fuel
plus air while  the specific  volume  of the base  case  (VB)  is
for air  alone.   Therefore,  the mass  flow  change includes the
change  in  the  mass  of  air,  and  the change  in the  mass  of
fuel.*   In  order  to   identify   the  change  in  volumetric
efficiency, we  must  separate  out the  change in  the  mass  of
the  intake air  from  the  total  mass  change.   This  can  be
determined  essentially  on  a  unit  ratio  basis  where  the
original air mass  is  unity, and the  portion  of  the  change  in
air mass  (M^c)  is t^16  ratio  of  the mass  of air (a)  to the
sum of  the mass of  air   (a)  plus the  mass  of  fuel  that has
evaporated (xf).

(AIII-12) MAC = (Mc/MB) (a/a + xf)

or,

(AIII-13) MAC = (Mc/MB)[l/(l + xf/a)]

Since we considered  the  old  air  mass equal  to  unity,  we can
now  compute  the  change   in air  mass  (CAM)  between  the  two
conditions .
*The change in mass of  fuel  refers  to the gaseous portion and
hence  is  essentially  governed  by  the   fraction  evaporated.
Liquid fuel is neglected.

-------
                             -37-

 (AIII-14) CAM = MAC - 1

or,

 (AIII-15) CAM = (vB/vc)(vc/vB)0.5[a/(a + xf)] - 1

We can  use  this  change in mass  flow  to  compute  the change in
volume  flow at the  inlet  to  the  system (i.e.,  at the air flow
meter).  Recognizing  that the  change in mass  flow is uniform
across  the  system,  the change  at the inlet  can  be written as
a  ratio of  the  difference between the change  in  air flow due
to cooling  (MAc), and the baseline case, or

 (Ain-16) CAM = (MAC - MB)/MB

Using the  continuity  Equation  (AIII-7)  and  inlet conditions,
we can  rewrite AIII-16 as,

 (AIII-17) CAM = [(dVA)ACi -  (dVA)Bi]/[dVA]Bi

At the  inlet  to  the system,  the  temperature and pressure for
all practical purposes  is the  same between  the  charge cooled
case,  and   the  baseline  case  because  all  of  the  potential
charge  cooling  would  occur  farther  downstream.    Hence  the
inlet density for  the charged  cooled case  is  essentially the
same, and  cancels  in  AIII-17  leaving only  the  velocity term
 (V) and the area term  (A).  The product of the  velocity (V)
and area (A)  happens  to be the volume  flow  rate  (Q).   The
result  of  substituting this  identity into AIII-17 provide us
with a  change in volume flow rate

 (AIII-18) CAM = (QACi - QBi)/Qfii = CQi

If we define  volumetric efficiency (VE) as  inlet  volume flow
rate  (Qi)  divided  by  engine displacement per unit  time,  the
change  in  volumetric  efficiency  (CVE) would be  equivalent to
the change in volume flow rate (CQi).

(AIII-19) VE = Qi/(Displ./Time)

(AIII-20) CVE = CQi

(AIII-21) CVE = CAM

Substituting  the  conditions  resulting   from  any  charge-air
cooling  for  CAM  (AIII-15)   into  equation  AIII  -  21, we  now
have a  function for the  change  in volumetric efficiency (CVE)
as  a  function  of  the  specific  volume resulting  from  the
temperature change due  to fuel evaporation,  and  the amount of
fuel that has evaporated.

(AIII-22) CVE = (vB/vc)(vc/vB)-5[a/(a + xf)] - 1

-------
                             -38-

                         Appendix IV

             Observed Volumetric Efficiency Loss

The volumetric efficiency  (VE)  change is based on  a constant
pressure  drop  across  the  throttle  (i.e.,  constant  manifold
vacuum, MV).   Analysis  of  the  test  results from Figure  3  in
Appendix VII indicate the following relationships.

    Baseline:    VE = -2.43 MV -I- 70.50    (RSQ = .96)
    Location 1:  VE = -2.57 MV + 71.17    (RSQ = 1.0)
    Location 2:  VE = -1.99 MV + 63.44    (RSQ = .77)

Substituting a range of manifold vacuums  from  10  to 15 inches
Hg, we arrived at the following tables:

                         Table AIV-1

          Volumetric Efficiency vs. Manifold Vacuum

     MV
    10
    12.5
    15
Baseline
34.05
40.13
46.2
Location 1
32.62
39.05
45.47
Location 2
33.15
38.13
43.10
                         Table AIV-2

          Volumetric Efficiency Changes (from Base)

     MV        CVE (%)-Location 1        CVE (%)-Location 2
    10               -1.6                      -6.7
    12.5             -2.7                      -5.0
    15               -4.2                      -2.6

    Average Loss      2.8                       4.1

-------
                             -39-

                         Appendix V

                  Heat Transfer Calculation
I .   Equation  for  correlation  coefficient   (C)  by  Sieder  and
    Tate (4)"


(AV-1)  (Tb2 - Tbl)(D)(Pr)*667(u)~-14/4(L)[(To - Tb)ln] = C

       where

       [(To - Tb)  In] = (A - B)/[ln(A/B)]

     A = Tol - Tbl
     B = To2 - Tb2
     Pr = (Cpu/k)  @ Tb
     u = (u @ Tb/u @ To)

II .   Boundary Conditions

     A.  Boundary Conditions for All Locations
                                     k)
                                     1 @
                                      60F (ref. 2)
         1. Air flow = 19.6 CFM = 1176.4 CFH
         2. k = .014 (approx. = air) (conduction
         3. Cp = .240 for air; = .245 for phi of
            a. Cp = (R/J)[k/k-l]
            b. R = mR/m = 1544/m
            c. m = (% fuel) (phi) (32.043) + [!-(% fuel) (phi )] (28 .85 )
            d. k = 1.4 for air (ratio of specific heats)
            e. k = 1.38 for menthanol/air at phi = 1 (ref. 2)
         4. Geometry similar to tubular heater
         5. Re, Pr ,  Nu numbers computed at bulk temperature
            (Tb) = (.5) (Tbl + Tb2)
     B.   Boundary Conditions for Location 1

         1. For phi = 1
a)
b)
c)
d)
e)
f)
g)
h)
Tbl
Tb
ub
uo
Cp
Tol
To2
Tb2
= 19F
= 50F
= 3.7 (E-7)
= 3.9 (E-7)
= .240
= 80F
= 160F
= 80F


(32.17) (3600) Ibm/
(32.17) (3600) Ibm/






ft.hr
ft.hr




                                                        @ 50eF
                                                        @ 80F
            i) Re neglects fuel mass flow.

            For phi = .8
a)
b)
c)
Tbl = 31
Tb  =56
ub  = 3.7
                         (E-7) (32.17) (3600) Ibm/ft.hr @ 50F

-------
                        -40-

       d) uo  = 3.9 (E-7) (32.17) (3600) Ibm/ft.hr @ 80F
       e) Cp  = .240
       f) Tol = 808F
       g) To2 = 160F
       h) Tb2 = 80F
       i) Re neglects fuel mass flow

C.  Boundary Conditions for Location 2

    1. For phi = 1

       a) Tbl = 7.5F
       b) Tb2 = 908F
       c) Tb = 48.7F approx = 50F
       d) ub = 3.7 (E-7) (32.17) (3600) Ibm/ft.hr @ 50F
       e) uo = 3.95 (E-7) (32.17) (3600) Ibm/ft.hr @ 90F
       f) Tol = 180F
       g) To2 = 240F
       h) Re neglects fuel mass flow
       i) Cp = .241 for k est. @ 1.395

    2. For phi = .8

       a) Tbl = 8.4F
       b) Tb2 = 90F
       c) Tb = 49.2F approx = 50F
       d) ub = 3.7 (E-7) (32.17) (3600) Ibm/ft.hr @ 50F
       e) uo = 3.95 (E-7) (32.17) (3600) Ibm/ft.hr @ 90F
       f) Tol = 180F
       g) To2 = 240F
       h) Re neglects fuel mass flow
       i) Cp = .241 for k est. @ 1.395

-------
                             -41-

III. Calculations

     A.  Location 1, phi = 1

0  L/D = [(Tb2 - Tb)(Pr)-667(u)-0.14]/4(c)[(To - Tb)ln]

   _(Pr).667 = (Cp u/fc.).667 = C( .240) (3.7)(E-7) (32.17) (3600)/(0.

   - (u)--14 = (ub/uo)-0.14 =  (3.7/3.9)-0.14 = 1.0074

0  [(To - Tb)ln)] = (A - B)/[ln(A/B)] = 70.1

   - A = Tol - Tbl = 80-19 = 61

   - B = To2 - Tb2 = 160-80 = 80

8  Re = Dvd/u

   Where

         d = slugs/cu. ft. = 2.3(E-3) slug/cu. ft. @ 59F

         v = ft./sec.

         u = Ib.-sec./sq. ft.

         D = duct diameter.

0  For various estimates of intake duct hydraulic diameter

D (in.)   D  ft.)   A (sq.ft.)   v (ft.sec.)     Re
2
2.5
3.0
.167
.208
.25
.022
.034
.049
14.989
9.586
6.662
1.56(E+4)
1.28(E+4)
1.04(E+4)
.00375
.00380
.00395
    L = (D/C)(K)(Tb2 - Tbl)/4[(To - Tb)ln]

    - K = .8021

    For various estimates of intake duct hydraulic diameter

                        Estimate of L             L (hardware
    D (in.)         Required to Achieve Tb2        measurement)

      2.0"                    95.2"                    80"
      2.5"                   117.4"                    80"
      3.0"                   137.2                     80"

-------
                         -42-

B.   Location 1, phi .8

Pr, Re, C, and u are the same as in Case 1, phi = 1.0

[(To - Tb)ln] = (A - B)/[ln(A/B)] = 63.2

- A = Tol - Tbl = 80-31 = 49

- B = To2 - Tb2 = 160-80 = 80

For various estimates of intake duct hydraulic diameter

                    Estimate of L             L (hardware
D (in.)        Required to Achieve Tb2        measurement)

  2.0"                   84.8"                    80"
  2.5"                  104.6"                    80"
  3.0"                  122.3                     80"

C.   Location 2, phi = 1

D of intake track runner = 1.35 in. = .1125 ft.

Q per runner = Q/4 = (.327 ft.3/sec.)/4 =  .0818 ft.3/sec.

v = 8.22 ft./sec.

Re = dvD/u = 5.748 E+3

C = .004

pr.667 = .7376

(U)-.14 = 1.009

K = (Pr)-667 (U)-.14 = .7442

[(To - Tb)ln] = (A - B)/[ln(A/B)] = 161

- A = Tol - Tbl = 172.5

- B = T02 - Tb2 = 150

L = (D/C)(K)(Tb2 - Tbl)/4[To - Tb)ln]

L (calculations) = 32.2 in.

L (hardware measurement) = 14 in. (approx.)

-------
                         -43-




D.   Location 2, phi .8




Pr, Re, C, D, and u are the  same as in Case 2, phi  .8



[(To - Tb)ln] = (A - B)/[ln(A/B)] = 160.5




- A = Tol - Tbl = 171.6




- B = To2 - Tb2 = 150



L (calculated) = 31.9 in.



L (hardware measurement) = 14 in. (approx.)

-------
                             -44-

                         Appendix VI

      Changes in Pumping Work Due to Temperature Changes

It  is  generally accepted that if the  air-charge  of an engine
is  cooled,  the  volumetric  efficiency  is  increased.   The
question  is  then  what  is  the  general  magnitude  of  the
differences between the work to pump cold air versus hot air.

As  indicated in other  appendices,  if fuel  is  used to cool the
charge,  the volumetric  efficiency  may  or  may  not  improve.
For the  purposes of  this appendix, we will assume that there
is  no fuel  in   the  air charge,  and  that  the  air  charge is
cooled by some  means other than fuel evaporation.

The simple  model  we will use is  a long tube of constant cross
section  connected to  a  pump which  exits  to  the atmosphere.
In  order  to consider  only  the work  necessary to  pump a given
quantity  of fluid,  and not  the  work to speed  up the air, we
will  assume the  pump  exit  area  to be  the  same  as  the  long
inlet tube.  The  following conditions will be assumed:

    Position 1  (inlet)

    P! = Pa
    Vi = 0

    Position 2  (inside tube)

    V2 = 9.586  ft/sec  (from Section III,  Appendix V)
    D2 = 2.5 in = .208 ft (from Section III, Appendix V)

    Position 3  (Exit)

    P3 = Pa = PI
    V3 = V2
    D3 = D2

Equation AVI-1  (pgs.  213,  216,  Reference 4)  is a restatement
of  the Bernoulli equation,  and  in  effect  says that  the work
needed to  pump  a given  amount  of  fluid  across  our  simple
system from position  1  to  position  3 is  equal  to  the  work
needed to change  the velocity, plus  the  work  needed to change
the  hydraulic   head,   plus   the   work  needed   to  change  the
pressure, plus the work needed to overcome friction.

(AVI-1)  W =  (.5)V2 - g  h-(RT/M)ln(P3/P1)-(.5V2Lf)/R

The change  in   velocity  is  simply  the outlet  velocity minus
the inlet velocity.

-------
                             -45-

               (.5)V2 = .5(V3)2 - .

                   Vx = 0

                   V3 = V2

(AVI-2)        (.5)V2 = .5(V2)2

The change in hydraulic head is zero.

(AVI-3)     h = 0

The change in pressure is zero

                   Pi = P2

(AVI-4)    In (Pi/P2) = In 1 = 0

If we  look at  the change in work per  unit  length at position
2, the unit work  due to friction is

         = (.5)(V2)2(D(f)/(.5)(D2)

(AVI-5)  = (V2)2(f)/D2

Substituting AVI-2 to 5 into AVI-1,  provides  for  the work per
unit length to pump fluid through our simple system,

(AVI-6)  W = -(.5)(V2)2 - (V2)2(f)/D2

The friction factor (f) is related to the Reynold's No.

         Re = DVd/u
         d  = P/RT*
(AVI-7)  u  = (.3170)(E-10)(T)l-5[734.7/(T+216)]**

From Appendix II, we can  assume a  dry  air  temperature of 80F
at condition 2.   The lowest dew point  of  the  fuel air mixture
was around -12C  (10.4F).  Therefore, we will  assume that we
can cool  dry  air to  that  temperature  for  this  comparison.
Using these assumptions results in the following:


    Temperature   	u	      	d	      	Re

        80F      3.86(E-7)       2.214(E-3)      1.42(E+3)
      10.4F      3.46(E-7)       2.542(E-3)      1.82(E+3)
* d = mass density = w/g.
**T = R, from Reference (7).

-------
                             -46-

Since the  Reynold's  No.s  calculated are in the laminar range,
the friction factor  (f) becomes,  .

(AVI-8)  f = 16/Re

Therefore, the work  Equation  (AVI-6) becomes,

(AVI-9)  W = -(.5)(V2)2 -  (V2)2(16)/(D)(Re)

To determine the  magnitude of the difference in work per unit
length  to  pump  hot   air  (WH)  versus  cold  air  (We),  a
simple  ratio  can be   formed  with  Equation  AVI-9,  and  the
appropriate  values  for position 2 and  the  Reynold's  No.  can
be substituted.

(AVI-10) WC/WH =  Z

or,

(AVI-11) Wc =  .9785 WH

Because  the  velocity  (V2) and  area  (A2)  remained  the  same
between the  comparison of the required  pumping  work with hot
air  to  the  required   work  with  cold  air,   the   results  of
Equation   (AVI-11)    are   applicable    to  constant   volume
flow-rate.     In   this    case,    artifically   reducing   the
temperature  of dry  air  by approximately  70F,   reduced  the
estimated  work  per   unit  length  by   slightly  more  than  2
percent.

However,  power  is   a  function   of mass  of  air   flow,  not
volume.  If  we want  to hold  mass flow  (M)  constant for  this
exercise, then Equation AVI-13 tells us that mass density (d)
and velocity must vary  with temperature.


(AVI-13) M = dVA

Since  we  have  previously  calculated  the  change  in  mass
density (d) with  temperature,  we can determine the  subsequent
change  in  velocity  by  forming a ratio of the  cold mass  flow
rate to the hot mass flow  rate.  Cancelling terms, we have

(AVI-14) VC/VH = dH/dc  =  .87097

Substituting AVI-14  into  AVI-9 and 10  results  in  the ratio of
work  per  unit  length  to  pump   the  same  mass  flow  at  two
different temperatures.

(AVI-15) WC/WH =  .87097 (Z)

-------
                             -47-

or,

(AVI-16) We = .8522 WH

The results of AVI-16 indicate that  the  work  to  pump the same
mass of cold fluid is approximately  15 percent less  than that
required for a hot  fluid.   But remember that this is  for  air
that was  cooled  by  some  means other  than fuel  evaporation.
Also this was a very simple  example  that did  not consider  the
intricacies in the geometry  of a real engine.   The  potential
effects of  the practical  engine  cycle  (intake,  compression,
power,   exhaust,   and   valve  overlap)  were  similarly  not
considered.  And  finally  the  myriad  of effects  involved  in
pulsating flow were  ignored.  Nonetheless, the  general  trend
is  that  if you  can cool  the air  by some  means other than
adding  fuel,  it  will  require less  power  during the  intake
stroke.

The amount of  work  reduction is,  however,  obviously not only
dependent on the  percentage reduction,  but  also  on  the base
level  of  intake  work.    Furthermore,   the  relation  of  the
intake work to the  other  losses  in  the  engine cycle  must  be
of  sufficient  magnitude,  in  order  for  a  small  percentage
improvement  in intake  work  due  to  a  cooler   charge   to  be
measureable in the overall engine performance.*
*Note:enginepower  increases due  to a cooler  intake  charge
are  essentially  due to  an  increase  in  charge mass  density
which increases the mass throughput of the engine.

-------
             -48-
         Appendix VII
EPA MEMO:   Results of Methanol
           Fumigation Investigations

-------
                              -49-
I
                UNITED STATES ENVIRONMENTAL PROTECTION AGENCY

                             ANN ARBOR. MICHIGAN 48105
   DATE: ADD   iqnq
         nrn C C. 1JO\I

SUBJECT: Results of Methanol Fumigation Investigations
                                                        OFFICE OF
                                                   A(R NOISE AND RADIATION
PROM
     TO
   THRU
         R. Bruce Michael
         Technical Support Staff
William Clemmens/ Project Manager
Technical Support Staff

Charles L. Gray, Jr., Director
Emission Control Technology Division
Phil Lorang, Chief
Technical Support Staf
                                      <
 The  hypothesis  of this study was: "Can methanol  injected  into
 the  upstream air passages  (additionally to  the normal  port
 injection)  cool the inlet  air  sufficiently through  vaporiza-
 tion to  increase  the  volumetric efficiency,  and  will  the
 expected  increase  in  volumetric  efficiency  translate  into
 improved  thermal  efficiency?"   This  hypothesis  is  derived
 from the mathematical equation  of volumetric  efficiency  in
 which  the efficiency  is  proportional  to the  mass  of air  per
 unit of  time flowing through the engine relative to  the swept
 volume of the engine during that time.  Vaporization (through
 fumigation)  in  the inlet air should  cause  a  temperature  drop
 of  the air, which would increase density  and mass  flow.   To
 test this hypothesis, two  methods  of  fumigation were  tested,
 both occurring along  with  the  normal port  injection  of  the
 Nissan  2.0  litre  NAPS-Z  engine.   Neither  method  improved
 volumetric  or thermal efficiency.   In fact,  the normal  port
 injection without  fumigation  gave  slightly  better  results.
 Because  none of  the  fumigation  systems  tested demonstrated
 any  increase in efficiency at  the  lower  power levels  tested,
 we  recommend that we abandon any plans for  fumigation  testing
 on  this engine  in the  future.

 Test Configuration

 Fumigation  of methanol was obtained by  adding EFI  style  fuel
 injectors  upstream  of   the  standard port  injectors.   The
 fumigation  injectors were  controlled  by the  EFI computer  in

-------
                            -50-


the  same  manner  and  simultaneously  with the  standard  port
injectors.  The  standard  port  injectors are located  approxi-
mately 3 to 5 inches from the intake valve seat.

Three  locations  for  the  fumigation  injectors  were  tested.
The first location was in  the  intake  manifold collector.   The
collector is  a log style manifold downstream  of the  single
throttle valve.  Four intake  runners  enter the bottom  of  the
collector.  They are ranged  in pairs  -  cylinder 1 and  2,  and
cylinders 3 and 4.  The intake  runners  are on the order of 12
inches  from the  intake  valve seat  to  the  collector.   The
single fumigation injector was mounted  on the top  side  of  the
collector,  and  roughly   equally  spaced  between  the  intake
runner  pairs  of  cylinders 1  and  2  and  cylinders   3  and  4
(i.e., between cylinders 2 and 3).

The  second  location  was also   in  the  collector,   but  two
injectors were  used instead  of a  single  injector.   The  two
injectors were  located over  the  intake  runner  pairs  -  one
injector over the pair  of runners for cylinders 1 and  2,  and
the other  injector  over  the pair  of  runners for  cylinders  3
and 4.

The third location utilized a  single  injector placed upstream
of the entrance to the EFI vehicle  system.   This  injector was
centrally located  in  a short  section  of  4  inch  O.D.  tubing
connected  to  the downstream   end  of  the  intake  air  flow
meter.    This   configuration   is   identified   as   "central
injection" or "central fumigation".

Engine Operation
                                  i
Engine  operation with  fumigation  eminating from the  first
location was not satisfactory.  Rough engine running, extreme
sensitivity  to  ignition  timing  and  ever present detonation
were characteristics of this location.  We speculate that the
central location of the injector  created  a situation in which
cylinders 2 and 3 robbed  fuel  from  cylinders  1  and 4 creating
a very lean condition in  cylinders  1  and  4.   Because of these
problems,  all  fumigation  testing   at  this   location   was
terminated.

The second location for the fumigation  injectors  seemed (to a
degree) to solve  the  assumed  distribution problems,  but  on  a
qualitative basis  the  engine  still  did not  operate  quite as
well as  in  the standard  configuration  (i.e., no  fumigation).
For  instance,  some of  the tests at  90  ft.lb.  and  2000  RPM
were scrubbed from the test program because  of  an instability
in the  MBT  point for  ignition  timing,  which led  to creeping
detonation.  Another handicap with  the  two injectors was that
even with  the  adjustable  air/fuel  control  box  for   the  EFI,
the sum  of  the  fuel  flow from  the two  fumigation  injectors

-------
                            -51-

plus the four standard injectors made it difficult  to  achieve
equivalence ratios leaner than 0.8.  Operation  at  equivalence
ratios  richer  than  0.8  would  normally  depress  the  brake
thermal efficiency somewhat, but Table 1 and Figure  1  suggest
that peak efficiency with the two injector set-up  is slightly
richer than  0.8.   Directionally the shift  to  a richer  point
from the baseline  (0.7 to  0.8)  for  peak efficiency  is in  the
wrong direction.

Only the  central  injector  seemed  to operate as  well as  the
standard configurations.   No difference  in driveability  was
observed.  The only handicap encountered was that  the  central
fumigation system  was  potentially  capable  of   running  leaner
than our adjustable EFI control box  would allow.

Results and Conclusions

The most interesting result of this  testing was that upstream
injection (fumigation)  of methanol  did not  improve volumetric
efficiency.  In fact,  as  shown  in  Table 1 (and Figures  1  and
2),  fumigation of methanol  on  this  port  injected  engine
actually decreased volumetric  efficiency by 4  to  10 percent.
In  three  out  of  four  cases,  the  decrease  in  volumetric
efficiency   resulted   in   a   noticeable   loss   in  thermal
efficiency (5-9%).  The only  case  that showed  an  increase in
thermal efficiency (0.5 percentage  point), with a  decrease in
volumetric efficiency was  the  test  point at high  power  using
the central  upstream  injector (Inlet  to EFI,   Table 1).   The
trend  of  this  increase  in efficiency  as shown  in Figure  2
(central  fumigation)   is  plateauing  around   an  equivalence
ratio of 0.8,  the  limit  of our  EFI control with  the  central
injector.   Potentially,   improvement  could  be  obtained  at
leaner equivalence ratios,  but  returning to Figure  1, we  see
that we were  able  to sufficiently  enlean the system to  get  a
peak in  the  efficiency trend with  the central  injector.   At
the  lower  power  in  Figure  1,   the peak efficiency for  the
central  injector   was   below  the   peak  efficiency  for  the
baseline.

It  is  not  known  exactly  why   fumigation  did  not  increase
volumetric efficiency,  but  information from the recent Nissan
work may shed  some light.   Nissan  was surprised to  find that
efficiency went down with  methanol, as  compared  to gasoline,
and  theorized that fuel  vaporization  was  caused  mainly  by
absorbing heat  from  the  cylinder walls  rather  than the air.
This  would,   of course,   result  in  little  or  no  gain  in
charging efficiency since  the incoming  air  would  not  change
density.   The lack of  a  density change results  in  a negative
effect  on  volumetric  efficiency for  the  following  reason.
When the fuel  vaporizes,  it increases  the  gas  volume  in  the
air stream and  has the effect of restricting or  slowing down
the  air  flow.  This  lowers  charging  efficiency,   more  than

-------
                            -52-

offsetting any gain.  While  this is not exactly  analogous  to
our fumigation work,  it may  be that  the  restriction of  air
flow  (due  to  fumigation)  was  the  dominant  effect,  which
lowered the volumetric efficiency.

-------
                              -53-

                           Table 1

                     Methanol Fumigation
                        Test  Results
                                         Injector Position
                                  Intake       Intake     Inlet to
                                 Collector    Collector      EPI
                       Baseline* 1 Injector  2 Injectors 1 Injector
90 ft-lbs @ 2000 RPM

Best Average Efficiency  38.9
Equivalence Ratio          .79
Volumetric Efficiency    69.3
Number of Tests           9

29.5 ft-lbs @ 1500 RPM

Best Average Efficiency  26.4
Equivalence Ratio          .70
Volumetric Efficiency    38.0
            _**
35.5
  .93
65.5
3
                       24.4
                         .86
                       33.8
39.4
  .83
65.1
3
            25.2
              .78
            35.5
Number of Tests
12
*Based on replicate  tests  before  fumigation testing (V87,V90)
and  replicates  after fumigation  (ZM2-VO).   The  best  average
(thermal)  efficiencies  were  calculated  as follows.   First,
the  data  were   separated   into  different  equivalence  ratio
ranges   for   each   version.   Second,  the   average  thermal
efficiencies  of  the  tests  in  each range  for each version was
calculated.   Third,  for  the equivalence  ratio  range yielding
the  overall  best efficiencies,  the average efficiencies for
the  three  versions  were  averaged.   For  the  higher  torque
level, the  equivalence  ratio  range which yielded the highest
average  efficiency was  .77-.80.   For the lower  torque level,
the range was .67-.73.

**A11 testing was aborted  in  this  configuration  due to severe
detonation problems  (no data taken) .

-------
                         -54-
  1500BPM
35
30
20


'







6*i*.;w*
'-A^-J^M* i 	 _/"
-* ** 	 + ' 	 r-4-K_
	 i 	 i 	 i 	 1  	 1 	
X BASELINE
N-15
 TWO OUTSIDE
INJECTORS
N-3
* CEMTRfiL INJECTOR
N-6
0 DISASSOCIRTION
BASELINE
N-9




 .65    .70   .75    .80
                     PHI
.85   .90
.35
              Methanol Fumigation
                   Figure 1

-------
                          -55-
    2000RPM
z
UJ
40

35
30
25
20
.E
c*7(\A<-
Far*i(.AriOtJ
")4ti|T)hr->^ '
G^Sii'-'vtf A.
* C.ENTRRL INJECTOR
NS
0 -DISRSSOCIRTION
BftSELINE
N-9


5 .70 .75 .30 .85 .90 .95
PHI
                Methanol Fumigation
                     Figure 2

-------
                            -56-

Purther  analysis  of  the  data  was  performed   to   compare
volumetric  efficiency  to  manifold  air  pressure  (manifold
vacuum).   Theoretically,   if   the  upstream  fumigation   of
methanol  cooled  the  inlet  air,  the  volumetric  efficiency
should increase for the same manifold air pressure.

Figure  3  shows   scatter  plots  and   regression  lines   of
volumetric efficiency versus manifold  pressure (MAP)   for  the
normal  port  injection  (baseline)  and  the  two  fumigation
versions.   The   regression  lines,   which  are   all   nearly
parallel, again show no advantage to fumigation.*

Figure  4  shows  thermal  efficiency  (BTE)  versus  manifold
pressure for the  same  tests.   In all cases, for  a  given  MAP,
efficiency was higher  in  the  baseline configuration,  thus
suggesting   that   methanol   fumigation   does   not   improve
volumetric efficiency.
*We should mention  that  the MAP measurements were  assumed to
be "dry" measurements  (no partial  pressure  of the  fuel in the
manifold), which is not  really  correct.   The fumigation would
really  result  in a  somewhat wet  condition in  the manifold,
making  the measured  MAP  a  wet measurement  which is by nature
greater  than a  dry measurement.   Therefore,   the  fumigation
points  and regression  lines really should  be slightly to the
left of those  shown in  Figure  3 which  translates  into worse
results.   No  calculations   were  made  to  account  for  this,
because the effect is slight.

-------
o
z
UJ
*
CJ
52


50


48


46


44


42


UO


38


36


34


32


30
    VOL.  EFF.   VS   MflP
    6
                               _L
          10   11   12   13   14

              MANIFOLD VACUUM
15   16
        6  BflSELINE.  TQ=29.5
        N=33     RSQ=.96
        60=70.50 Bl-2.43

        +  TWO INJECTORS
        N=ll     RSQ=.77
        80=63.44 B1=-1.99

          1  CENTRflL
        INJECTOR
        N=7       RSQ=1.0
        80 = 71 .17 Bl=-2.57
                               Figure 3

-------
  28
27
u 26
  25
cr
a:
cc
cc
  22
    8
         BTE VS MflP
                                       +
                                        +
            10    11    12    13

                 MANIFOLD VACUUM
14
15
                                             BflSELINE, TQ=29.5
                                            N = 33
                                            + TWO  INJECTORS
                                            N=l 1
                                             CENTRflL INJECTOR
                                                                        Ul
                                                                        oo
                               Figure 4

-------
                            -59-

Prom  another aspect,  graphs  of  BTE  vs.  MAP  can  also  be
valuable.  The  air-fuel  ratio  was varied from an  equivalence
ratio (PHI)  of  about  .6  to .8  for many test points.   For  the
power output and rpm to  remain constant, the throttle opening
has  to   be  varied  somewhat  inversely  to  the  different  PHI
levels - a  lower  PHI  requires  a larger throttle opening,  for
example, which  in turn  lowers  the throttling  (pumping)  loss
because  of  less restriction of  the  air  flow.   Therefore,  it
can be determined from the results if  throttling loss changes
have an  appreciable  effect on  thermal efficiency  when leaner
mixtures (leaner than .8) are used.

Results  from baseline  tests show  that leaner  mixtures  (with
lower throttling  losses)  did  not  improve thermal  efficiency;
in fact  the  reverse  was  true.   Figures 5-8 show baseline  (no
fumigation)  test  data  for four  categories  of  PHI.   Figure  5
includes test  points  at the  lowest  levels of  PHI  and  MAP,
which produced  the  lowest  levels of  BTE.  Figure  6,  which
includes intermediate  low  values  of  PHI  and MAP,  shows  mixed
results  ranging from nearly  the lowest to the highest levels
of  BTE.   PHI   values  in  this  category  appear  to  be  the
transition from low to high BTE.   The  highest  BTE  values were
from  PHI at  .66-.67   and  the  higher  MAP  values,  while  the
lowest  BTE  values were  from  PHI  at  .63 and  the   lowest  MAP
values.   Figures 7 and 8 show  consistently high  levels of  BTE
for PHI  greater than  .67  and  MAP greater  than  13  inches  of
mercury.   Thus  it  appears  that  throttling   losses  are  not
dominant characteristics  on this engine under  these loads.

In summary,  fumigation did not  improve  either volumetric  or
thermal efficiency and we therefore recommend  not  pursuing it
further on this engine.

-------
BTE VS MflP  BflSELINE   .58
-------
BTE VS MflP   BflSELINE   .63
-------
BTE VS  MflP   BflSELINE   .68
-------
      -63-
 Appendix VIII

   Test Data
(Summary Sheets)

-------
TEST NO.: HO-81O809
                          ENGINE:  38O NAPS-ZM
                                                      87
          TEST 0/T:  B-31-82  O:, O
               REPORT D/T: O9-27-B3  13:57
                                     PO625B1
SUMMARY REPORT
     TORO. MAX THR  IGN.  INJ.  CORR.
RJ>L FT-LB TO. POSN TIMG. T1MG.  BHP
                                34.91
                                35. 12
                                34 .74
                                11.41
                                 8 67
                                 8.67
200O
2OOO
2OOO
15OO
15OO
1500
90
91
90
39.
30
30.
.5
.O
.O
5
0
O
100
1O1
99
10O
76
76
2 1O
110
oao
1270
13<4O
1O70
22
23
25
32
33
37
.OB
.SB
.SB
.SB
.SB
.OB
O.
O.
0.
O.
O.
O.


MEAS.
A>
7
8
8
a
8
9
If
.75
.21
.42
O3
.87
.80
PHI
EOUIV
RATIJJ
O.83
0 79
0.77
OBI
0 73
0.66
                   BSFC
                   GAS
             BSFC  EOUIV
       FUEL  LB/   LB/
       LB/HR BHPHR BHPHR
  ENERGY
EFFICIENCY
     1000*
     MBTU/
 %	 BHPHR
206 7  26.66 O.764 0.359 38.6
217.9  26.54 O.756 0.355 39.O
22O.2  26.14 0.753 O.354 39.1
 9O.8  11.31 O.991 0.466 29.7
 87 6   9.88 1.14O O.536 25.8
      6.60
      6.53
      6.51
      8.57
      9.85
	 BRAKE SPECIFIC EMISSIONS (G/BHPHR)  	
 BSHC   BSCO    BSC02   BSNOX BSALDYH BSPART
 1 .62O
 1.878
 1 .895
 2.707
 6.33O
                                                    1O3 3   1O.55  1.216 0.571 24.2  1O.52  14.888
1.388  408.16  9.555
2.537  398.25  7.SOS
2.S6O  398.O1  6.318
3.427  56O.25 11.878
5.272  597.08  4.804
6.761  473.22  1.673
O.O
O.O
O.O
0.0
O.O
O.O

-------
TEST NO.:  HO-81O81O
                         ENGINE: 38O NAPS-ZM
                                                     87
                                   TEST D/T:  9-  1-82  8:3O
                                 REPORT  D/T:  O9-27-83 13:56
                                                                                                                              PO625B1
SUMMARY REPORT
     TORO. MAX THR  IGN.
RPM  FT-LB TQ. POSN TIMG.
2OOO
2OOO
2OOO
150O
15OO
1500
91
90
89
29
29
28
.O
.0
.7
.5
5
0
IOO
99
99
IOO
IOO
95
47O
1 IO
080
146O
1310
1060
22
25
25
33
31
34
OB
. 58
OB
. 8B
OB
2B
O.
O.
0.
0.
O.
O.
                   PHI
INJ.   CORR. MFAS.  EOUIV
TIMG.   BHP   A/F   RAMO
            BSFC
            GAS
      BSFC  EOUIV
FUEL  LB/   LB/
  /HR BHPHR BHPHR
  ENERGV
EFFICIENCY
     1OOO*
     MBTU/
 %   BHPHR
	 BRAKE SPECIFIC EMISSIONS  (G/BHPHR) 	
 BSHC   BSCO    BSC02   BSNOX BSALOYH BSPART
                                35 43  7.71  O 84  2O5 8  26.7O O.754 0.354 39.1  6.52   1.961    2.677  441.49  9.687
                                                          26.12 O.745 0.35O 39.6
                                                          26.36 O.754 0.354 39.0
                                                           9.55 I.1O9 O.521 26.6
                                                           9.83 1.142 0.537 25.8
                                                          10.32 1 .267 O.595 23.2
O.
0.
0.
O.
O.
35
34
a
a
8
07
94
.GO
.60
. 15
8
8
8
8
10
26
.23
.38
92
O1
0
O
O
O
0
78
79
77
73
G5
215
217
80
87
IO3
. 7
.O
O
6
3
6
6
9
9.
IO.
44
,52
59
87
95
2
2
8
a
14
.384
.599
.597
.588
.545
2
2
4
5
7
.510
.515
.495
.394
. 6O1
433
436
618
614
623.
.92
.46
.02
77
48
8
7
8
3.
1 .
.678
.837
. 3O5
911
312
                                                                                                      0.0
                                                                                                      O.O
                                                                                                      O.O
                                                                                                      O.O
                                                                                                      O.O
                                                                                                      O.O
                                                                                                                                        U1
                                                                                                                                         I

-------
TEST NO.: HO-8IOS58
                    ENGINE:  380 NAPS-ZM
                                                      90
                                                  TEST 0/T:  9-21-82  9:45
                                                                                            REPORT O/T: O9-27-83  13:56
                                                                                                                        P062581
SUMMARY REPORT
                                                                       BSFC    ENERGV
                                                                       GAS   EFFICIENCY
            %                                 PHI                 BSFC  EOUIV      1OOO*
     TORQ. MAX THR  IGN.   INJ.  CORR.  MEAS.  EQUIV   AIR    FUEL  LB/   LB/        MBTU/
RPM  FT-LB TQ. POSN JJ.MG_._  T IMG . _BHP_   A/F   RAT_m  _LB/HR  LB/HR BHPHR BHPHR  %   BHPHR
                                                                                  	 BRAKE SPECIFIC EMISSIONS  (G/BHPHR) 	
                                                                                   BSHC   BSCO    BSC02   BSNOX BSALOVH BSPART
20OO
15OO
9O.5
29.5
1OO
1OO
 110 24.OB
I3OO 38.OB
0.
O.
35. 10
 8 54
7 .88
9. 16
O.82
0.7 I
208 9
 89 9
26.51
 9.81
755 0.355 39.O
148 0.539 25.7
6.53
9.93
1 .608
5.231
2. 118
4 . 3O4
439.1O
605.85
8. 058
3.511
O.O
O.O
                                                                                                                                        I
                                                                                                                                        a\

-------
TEST NO.: HD-8IO859
                          ENGINE:  3BO  NAPS-ZM
                                                      90
                                                  TEST D/T:   9-21-82  O:  O
                                                                             REPORT D/T: O9-27-83  13:55
                                                                                                                               PO62581
SUMMARY REPORT
            y.                                 PHI
     TORQ. MAX TIIR   IGN    1NJ.  CORR.  MEAS.  EOUIV   AIR
RPM  FT-LB TQ. POSN  T1MG.  TIMG.  JHP_  A/F   RAM0   LB/HR
2OOO  90.O IOO   I4O 24.OB
2OOO  90.5 1O1
2OOO  89.5
150O  29.5
15OO  29.5
150O  29.5
 99
IOO
too
IOO
  9O 24 OB
  80 27.OB
145O 36.OB
1'JO  4 2 . OB
 91O 48.OB
                            O.  34 65  7 82  O  83   2O5  8
                            O.  34.83  7 96  O  81   2103
O.
O.
O.
O.
34 .45
 851
 8.51
 851
 8. 17
 8.56
 9.42
10. 3O
O.79  2112
O.76   80 4
O.69   89 9
O.63  1O9 6
                                                           BSFC     ENERGY
                                                           GAS    EFFICIENCY
                                                     BSFC  EOUIV      1OOO*
                                               FUEL  LB/   LB/         MBTU/
                                               LB/HR BHPHR BHPHR  %   BHPHR
                                                                               BRAKE SPECIFIC  EMISSIONS (G/BHPHR)  	
                                                                          BSHC   BSCO    BSC02    BSNOX  BSALDYH BSPART
6
6
6.
9
9
.57
56
49
54
.69
1.
1
1
4
5
.346
. 8O6
9O4
1-19
439
2
2
2
3
4
. 149
.262
.247
259
.057
335
437
433
6OO.
596
.40
. 19
69
13
. 13
a
6
6
4 .
1
.523
.777
.662
.909
.498
O
O
O.
O
O.
O
.0
.0
0
O
26.33 O.76O O.357 38.8
26.42 O.759 O.356 38.8
25.85 O.75O O.353 39.2
 9.39 1.1O4 O.519 26.7
 9.54 1.121 O.527 26.3
1O.64 1.250 O.587 23.6  1O.BO 17.268   8.327  594.76  O.91O
                                                                                                                                  O.O
                                                                                                                                          I
                                                                                                                                         
-------
TEST NO.:  HD-8IO86O
ENGINE:  3BO NAPS-2M
                                                     90
                                                             TEST D/T:   9-22-82   8:45
REPORT D/T: O9-27-83 13:55
P06258t
SUMMARY REPORT



RPM
2OOO
2OOO
2OOO
15OO
15OO
1500


TORO.
FT-LB
90. 0
90.0
89.3
29.6
29.5
30.8

%
MAX
10^
1OO
100
99
1OO
1OO
1O4


THR
POSN
12
O9
O8
142
125
95


IGN.
T1MG.
22. OB
22.08
24 .OB
44. OB
47. OB
47. OB


1NJ.
TIMG.
O.
O.
0.
O.
O.
O.


CORR.
BHP
34 .88
34. 9O
34 5O
8 56
8.54
8 92


MEAS.
A/F
7 .79
7 .80
7 96
8.51
9. 34
1O. 19

PHI
EOUIV
RA1 10
O.83
0 83
O 81
0 76
O.69
O.64


AIR
LB/
208
210
211
83
92
IO9



HR
.O
.3
.2
1
.6
2


FUEL
LB/HR
26.70
26.95
26.53
9.77
9.91
1O. 72

BSFC
LB/
BHPHR
O.765
O.772
O.769
1 . 142
1 . 161
1 .202
BSFC
GAS
EOUIV
LB/
BHPHR
0.36O
O.363
0.361
O.536
O.545
0.565
ENERGY
EFFICIENCY


%
38.5
38. 1
38.3
25.8
25.4
24.5
1OOO
MBTU/
BHPHR
6.62
6.68
6.65
9.87
1O.O3
1O.39



BSHC
1.351
1.411
1.542
16.7O4
7.O44
17.257



BRAKE SPECIFIC EMISSIONS
BSCO
1 .823
1 .932
1 .987
3. 195
4.57O
6.97O
BSC02
451 . 18
449.84
446.63
621 .66
613.92
578.51
BSNOX
9. 1O7
7.654
8.247
9.857
4.772
1 .592

(G/BHPHR) 	
BSALDYH BSPART
O.O
0.0
0.0
O.O
0.0
O.O
                                                                                                                                       CO
                                                                                                                                       I

-------
TEST
NO. : HO-81O861
SUMMARY REPORT
'/,
TORQ. MAX
RPM
2OOO
2OOO
2OOO
2OOO
15OO
I5OO
1501
15OO
FT-LB
9O.O
89.5
89.5
89.5
29.5
30.5
29.5
29.5
TO.
too
99
99
99
1OO
103
1OO
100
THR
POSN
I2O
too
O9O
O8O
143O
13-1O
1 170
9-40
ENGINE : 38O NAPS-ZM
IGN. 1NJ.
TIMG. T1MG.
21
23
22
22
38
42
55
56
.OB
OB
.88
BE
OB
.58
5B
.58
O.
O.
O.
O.
O.
O.
0.
O.
CORR.
BMP
34
34
34
34
8
a
8
8
.at
62
61
61
.56
84
56
56
MEAS.
A/F
7.61
771
7.73
7 57
8.O7
8.42
9.26
0.0
PHI
ECU IV
RATIO
O
o
o
o
0
0
o.
o
85
84
84
85
80
77
7O
O
9O
AIR
TEST
0/T:
BSFC
FUEL LB/
LB/HR
204
2OG
206
204
78
83
93
O
4
. 7
7
4
6
1
.5
O
LB/
26
26
26
26
9
9
IO
IO
'HR BHPHR
87 O.
. 80 O.
.75 O.
99 O.
74 1 .
.87 1 .
09 1 .
.64 1 .
772
774
773
7 BO
139
1 16
178
243
: 9-22-82 13:45
BSFC
GAS
EOUIV
LB/
ENERGV
EFFICIENCY
1OOO*
MBTU/
BHPHR %
0.
0.
0.
O.
O.
0.
O.
O.
363
364
363
366
535
524
553
584
38
38
38
37
25
26
25
23
.2
.O
. 1
.8
.9
.4
.O
.7
BHPHR
6
6
6
6
9
9
1O
IO
67
.69
.68
74
.84
65
. 18
.74


REPORT


BSHC
1
1
1
1
2
3
a
19
.462
.583
.623
.637
.882
. 58O
. 55O
.079
D/T:
O9-27-83
13:54
PO625B1
JAKE SPECIFIC EMISSIONS (G/BHPHR) 	
BSCO
1
2
2
2
3
3
5
7
.912
.014
. 1O2
. 1 14
. O91
.433
.355
.4 IO
BSC02
455
45O
451
455
624
604
609.
593.
.68
.69
.36
.56
.25
.45
22
28
BSNOX BSALDVH BSPART
9
8
8
8
9
7
4 .
2.
.348
.974
. 194
200
.696
.237
131
O66
O.O
O.O
O.O
0.0
O.O
0.0
00
O.O

-------
TEST NO.: HO-B1OB62
                          ENGINE:  38O  NAPS-ZM
                                                      91
                                                        TEST  D/T:  9-22-82  16:30
                                                                            REPORT  D/T:  O9-27-83 13:48
                                                                                                                             PO62581
SUMMARY REPORT
            X                                 PHI
     TORQ. MAX THR   IGN.   INJ.  COPR. MEAS.  EOUIV
RPM  FT-LB TO. POSN  TIMG.  T1MG.  BMP   A/F   RATIO
2OOO
2OOO
2OOO
15OO
150O
9O.O 1OO
91.O 1O1
89.5  99
1 1
O8
OS
29.O 1OO 145
29.O 1OO 134
17. 5B
16.OB
17 .SB
39. 5B
4O.8B
15OO  3O.O 1O3 1180 42.2B
                                                                 BSFC
                                                                 GAS
                                                          BSFC   EOUIV
                                                     FUEL  LB/    LB/
                                                     LB/HR BHPHR  BHPHR
                                                              ENERGY
                                                            EFFICIENCY
                                                                  1OOO*
                                                                  MBTU/
                                                             %    BHPHR
                                                                    	 BRAKE  SPECIFIC EMISSIONS (G/BHPHR)  --
                                                                    BSHC   BSCO    BSC02   BSNOX BSALDYH BSPART
                                 8 69  7.89  O 82
202 2  32.02 O.918 0.431 32.1  7.94  5.2O7
202 2  31.95 O.908 0.426 32.5  7.85  4.1O3
201 3  31.O4 0.896 0.421 32.9  7.74  4.796
 78.6  1O.98 1.3O4 O.613 22.6 11.27 15.957
 81 8  11.23 1.335 0.627 22.1 11.54 15.030
 92.1  11.67 1.342 O.631 21.9 11.6O  8.423
37.4O1
32. 175
25.769
55.446
44.581
471 .76
474.61
475.41
625.82
632.67
4.9O4
5.049
6.O97
9.828
9.25O
                                                                                                 5.276  6OO.16  4.O7O
0.0
O.O
O.O
O.O
O.O
O.O
                                                                                                                                          I
                                                                                                                                          -J
                                                                                                                                          O
                                                                                                                                          I

-------
TEST NO.: HO-S10863
                         ENGINE: 38O NAPS-ZM
                                                     91
                             TEST D/T:  9-23-82  9: O
                                                     REPORT D/T:  O9-27-83 13:49
                                                                                                                              PO62581
SUMMARY REPORT
     TORQ. MAX THR  1GN.  INJ.
RPM  FT-LB TO. POSN TIMG. 1IMG.
                            O.
                            O.
                            O.
                            O.
                            O.
                            O.
2OOO
2000
2000
1500
15OO
15OO
89
90
90
29
3O
29
.5
0
.2
.5
.2
.5
1OO
1O1
1O1
1OO
1O2
1OO
35O
210
14O
95O
146O
139O
2O
17
18
32
4O
4 1
.OB
OB
SB
.OB
58
8B
             PHI
CORR.  MEAS.  EQUIV
 BHP   A/F  RAI1O
34. 79
34 .83
34 .83
 8 55
 8.74
 8 55
            AIR
            LB^HR
6.58
7 . 16
7.56
  2O
  56
                                       8.23
O.98
0.9O
O 86
O.9O
O 8G
O 79
                         BSFC     ENERGY
                         GAS   EFFICIENCY
                   BSFC  EOUIV       1OOO
             FUEL  LB/   LB/         MBTU/
             LB/HR BHPHR BHPHR  %   BHPHR
                                          BRAKE  SPECIFIC EMISSIONS (G/BHPHR)  	
                                     BSHC    BSCO    BSC02   BSNOX BSALOYH BSPART
179 7  27.3O O.785 O.369 37.5  6.78   1.872
195 5  27.31 O.784 O.368 37.6  6.78   1.609
2O4 4  27.03 0.776 O.365 38.O  6.71   1.654
 71.0   9.86 1.154 O.542 25.5  9.97   5.228
 76.4  1O.11 1.156 O.543 25.5  9.99   4.242
 81 8   9.93 1.162 O.546 25.4  1O.O4   4.O99
4.982  46O.89  1O.ISO
1.485  459.36  8.972
1.599  457.28  9.109
3.944  628.36  10.944
2.912  635.69  14.O91
3.458  636.92  14.477
O.O
O.O
O.O
O.O
0.0
O.O

-------
TEST NO. : HD-81O864
SUMMARY REPORT
TORO. MAX TMR
RPM FT-LB
2OOO 9 1 . O
2OOO 87. O
20OO 88. O
150O 29. O
15OO 29.5
1500 3O.O
TO. POSN
1OO 35O
96 27O
97 1OO
1OO 152O
1O2 138O
1O3 1 1OO
ENGINE
IGN.
T1MG.
2O. OB
21 5B
24 OB
42 .OB
42 48
46. OB
1NJ.
T1MG.
0.
O.
O.
O.
0.
O.
: 38O NAPS-ZM
CORR.
BMP
35 19
33.64
34 .Ol
841
8 56
8.71
Ml AS
A/F
6 57
7 45
7 . 76
7 .47
8.39
9.41
PHI
EOUIV
RATIO
0 98
O 87
O.83
O 87
0.77
O.69
92
AIR
LB/HR
182 0
19 1 O
202 2
74 1
83 1
99 3
TEST 0/T
FUEL
LB/HR
27.69
25.63
26.07
9.92
9.91
IO.56
BSFC
LB/
BHPHR
O.787
O.762
O.767
1 . 18O
1 . 158
1.212
: 9-23-82
BSFC
GAS
EOUIV
LB/
BHPHR
0.370
0.358
0.36O
0.554
0.544
0.570
13: O
ENERGV
EFFICIENCY
1OOO*
MBTU/
/,
37 .4
38.7
38.4
25. O
25.4
24 .3
BHPHR
6.8O
6.59
6.63
1O.2O
1O.01
tO. 48
REPORT 0/T:


BSHC
O.692
O.995
1 .540
2.841
3.354
8.411
O9-27-83 13:5O PO6258I
RAKE SPECIFIC EMISSIONS (G/BHPHR) 	
BSCO
1 .474
1 .798
2.O69
2.971
3.453
5.268
BSC02
436.91
4 12 .89
413.55
579.60
565.09
599.31
BSNOX
12.993
7.277
4.735
12.O3O
4 . 162
4 . 064
BSALDYH BSPART
O.O
0.0
O.O
0.0
00
0.0
IVJ
 I

-------
TEST NO.: HD-8IOB65
                          ENGINE: 38O NAPS-ZM
                                                       92
                                                   TEST D/T:   9-23-82 14: O
                                                                              REPORT D/T:  O9-27-83 13:5O
                                                                                                                                PO62S81
SUMMARY REPORT
            y.                                 PHI
     TORQ. MAX  TIIR   IGN.   INJ.  CORR.  MEAS  COUIV   AIR
RPM  FT-LB TO.  POSN TIMG.  TIMG. _BHP_  A/F  RA11J) J.B/HR
2OOO  90.5  1OO   39O 18.OB
2OOO  BB.5  98   ?9O 21.OB
2OOO  90.O
150O  29.5
I5OO  28.8
15OO  31.O
 99
1OO
 98
105
 12O 25.5B
149O 40 58
138O 41.58
1200 45.BB
                 O.  35.O3   6.74   O.96  179.7
                 O.  34 25   7.34   O.8B  189 6
O.
O.
O.
O.
34 84
 8 56
 8 36
 9 OO
7 97
7 .44
8.37
9.3O
O.81  206.7
O 87   76 4
077   83  I
O.7O   94 4
FUEL
LB/HR
26.66
25.83
25.95
IO.26
9.93
IO. 15
BSFC
LB/
BHPHR
O.761
0.754
O.745
1 . 198
1 . 188
1 . 128
BSFC
GAS
EQUIV
LB/
BHPHR
O.358
O.354
O.35O
O.563
O.558
O.530
ENERGY
EFFICIENCY
1OOO*
MBTU/
_%_
38
39
39
24
24
26

. 7
. 1
.6
.6
.8
. 1
BHPHR
6.58
6.52
6.44
1O.36
1O.27
9.75


BSHC
O.8O9
0.977
1 .283
3.456
3.829
5. 100
RAKE SPECIFIC EMISSIONS (G/BHPHH) 	
BSCO
1
1
1
2
3
3
. 29O
.512
.764
.888
.483
965
BSC02
428.27
417.86
4O9.08
582.72
57O.88
502.53
BSNOX
1 1 .570
9.264
7.549
14 .407
4.O36
O.798
BSALOVH BSPART
O.O
0.0
O.O
0.0
0.0
O.O
                                                                                                                                             I
                                                                                                                                            -J

-------
TEST NO.:  HO-81O866
                         ENGINE:  38O NAPS-ZM
                                                     92
   TEST D/T:  9-23-82 15: O
               REPORT D/T: O9-27-83  13:51
                                                                                                                             PO6258 1
SUMMARY REPORT
     TORO. MAX THR  IGN.   INJ.  CORR. MEAS.
RPM  FT-LB TO. POSN TIMG.  TIMG.  BHP   A/F
                                             PHI
                                            EOUIV  AIR
                                            RATIO  LB/HR
            BSFC
            GAS
      BSFC  EOUIV
FUEL  LB/   LB/
LB/HR BHPHR BHPHR
  ENERGY
EFFICIENCY
     IOOO*
     MBTU/
 %   BHPHR
2OOO  89.O  98  3OO 21.58
20OO  9O.2  99  130 28.OB
1SOO  29.2 1OO 1490 41.OB
15OO  30.0 1O3 1410 49.8B
15OO  29.5 1O1 125O 49 2B
O.
O.
O.
0.
O.
34
34
8
8
a
41
.87
46
7O
56
7
8
7
a
9
.32
.08
52
. 1 1
.05
O
0
0
O.
O
88
BO
86
BO
7 1
188
2O7
76
81
9O.
7
6
4
a
a
     BRAKE SPECIFIC EMISSIONS  (G/BHPHR) 	
BSHC   BSCO    BSC02   BSNOX BSALDYH BSPART
2OOO  91.0 10O  42O 18.OB   O.  35.19  6.66  O.97   178 8  26.86 O.763 0.359 38.6  6.6O  O.851
                                                          25.77 0.749 0.352 39.3  6.47  O.9O1
                                                          25.70 0.737 0.346 4O.O  6.37  1.115
                                                          1O.16 1.2O1 0.564 24.5 1O.38  3.248
                                                          1O.O3 1.153 O.542 25.5  9.97  3.019
                                                          10.03 1.1720.551 25.1 1O.13  4.05O
                                      1.253  43O.36  12.169
                                      1.415  414.51  1O.553
                                      1.678  406.91  9.544
                                      2.817  568.25  12. 14O
                                      2.78O  553.85  7.2O4
                                      3.896  512.O6  1.124
                                                    0.0
                                                    O.O
                                                    O.O
                                                    O.O
                                                    O.O
                                                    O.O

-------
TEST NO.: HO-B1OBG7
                          ENGINE:  38O  NAPS-ZM
                                                      91
                                                  TEST D/T:   9-24-82 16: O
                                  REPORT D/T: O9-27-83  13:49
                                                                                                                               PO62SS1
SUMMARY REPORT
     TORO. MAX THR  IGN.  1NJ.
RPM  FT-LB TQ. POSN I IMG. TIMG .

15OO  3O.O 1OO 150O 4O.OB   O.
150O  29.5
15OO  29.5
15OO  29.O
15OO  29.5
15OO  29.8
98 I45O 44 .SB   O.
98 1350 24 .58   O.
97 154O 21.58   O.
98 145O 23 OB   O.
99 136O 53 58   O.
                                 PHI
                    CORR. MEAS. EOUIV   AIR
                     BMP   A/F  RAMO   LB/HR
            BSFC    ENERGV
            GAS   EFFICIENCY
      BSFC  EQUIV       1OOO*
FUEL  LB/   LB/        MBTU/
LB/HR BHPHR BHPHR  %   BHPMR
     BRAKE SPECIFIC  EMISSIONS (G/BHPHR)  	
BSHC   BSCO     BSC02   BSNOX BSALDYH BSPART
8. 74
8 6O
8 59
8. 44
8 59
8 .68
7. O2
7.68
8 6O
7.63
8. 23
8.57
O.92
O 84
0. 75
0.85
0 79
0.76
78 6
80 9
89 0
76 4
82 7
87 6
1 I .20
10.53
1O.34
1O.O1
10. 04
10.23
1 .281 O.6O2
1 .225 O.576
1 . 2O4 O.566
1 . 185 O.557
1 . 169 O.549
1 . 178 O.553
23.0 1 1 .07
24.0 1O.59
24 .5 1O. 4 1
24.9 1O. 25
25.2 1O. 1O
25. O 1O. 18
3.2O3
3. 1 14
4.228
2.269
2.320
2.966
13.291
5.575
4 . 066
1 1 .843
4.813
3.3O3
64O. 18
633. 10
633.52
615.45
61O.47
636.52
14 .966
12.415
6.504
4 .221
6.O49
1 1 .OO2
O.O
O.O
O.O
0.0
O.O
0.0
                                                                                                                                         Ul
                                                                                                                                          I

-------
TEST NO.:  HD-81O87t
                         ENGINE: 3BO NAPS-ZM
                                                     9O
TEST D/T:  9-27-82  9:  O
REPORT D/T: O9-27-83 13:53
PO62581
SUMMARY REPORT

RPM
2OOO
20OO
2OOO
15OO
15OO
15OO
2OOO
2017
2OOO
15OO
150O
1516
TORO.
FT-LB
89. 0
89.5
90. 0
29.5
3O.O
3O.O
89.5
B9.5
90.2
29. 0
3O.2
29. 0
y.
MAX
TO.
IOO
1O1
1O1
IOO
102
1O2
IOO
IOO
IOO
IOO
1O4
IOO
THR
POSN
12O
oao
oao
136O
1 13O
aao
120
O8O
oao
144O
129O
B7O
IGN.
INJ.
TIMG. TIMG
2O
22
22
44
50
46
22
21 .
21
28
44 .
52
5B
OB
5B
5B
OB
5B
OB
5B
OB
OB
SB
OB
0.
0.
O.
O.
O.
O.
O.
O.
O.
O.
O.
0.
CORR.
. BMP
34
34
35
8
8
a
34
35
35
8
8
8
80
87
09
62
77
77
89
17
14
47
83
56
MEAS.
A;
7
7
7
9
1O
10
7
a
7
a
9
10
if
.96
97
92
1O
O6
60
98
05
87
82
70
87
PHI
EOUIV
RATIO
O
0
0
0
O
O
O
0
o
o
o
o
.8 1
81
.82
.7 1
61
61
81
BO
82
73
67
6O
AIR
BSFC
FUEL LB/
LB/HR LBj
211
2 12
211
87
101
113
211.
215
211
81
9O.
1 16 .
2
1
2
6
t
2
2
7
2
a
a
a
26
26
26
9
1O
1O
26
26
26
9
9
1O.
fHR BHPHR
.51 0
.62 0
. 65 O
.63 1
,O5 1
69 1
48 0
81 O
, 84 O
27 1 .
36 1 .
75 1
.762
764
760
. 1 17
. 146
218
759
762
764
O94
O61
255
BSFC
GAS
EOUIV
LB/
ENERGY
EFFICIENCY
1OOO*
MBTU/
BHPHR %
O
O
O.
0
O
O.
0.
0.
O.
O.
0.
O.
358
359
357
525
538
572
357
358
359
514
498
59O
38.7
38 6
38.8
26.4
25.7
24 . 2
38.8
38.6
38.6
26.9
2/8
23.5
BHPHR
6
6
6
9.
9
1O.
6.
6.
6.
9.
9.
1O.
59
6O
57
65
91
53
56
59
60
45
17
85


Bi
1
1
1
O
1 t
19
2
t
2
2
2
2
 	 p(

5HC
.241
.541
.559
.O
.336
.589
108
983
O63
O63
OG3
.063
JAKE SPECIFIC EMISSIONS (G/BHPHR) 	
BSCO
2
2
2
O.
5
8.
2.
2.
2.
2.
2.
2 .
O5B
. 148
192
.O
768
538
161
24O
232
232
232
232
BSC02
441
397
437
O
566
559
445.
442.
444.
444.
444.
444.
.85
.81
.53
. 15
.38
37
25
91
74
74
74
74
BSNOX BSALOYH BSPART
7.O61
6.836
7. 167
o.ota
3.759
1 .332
7 .759
6.4O4
6.4O8
6.408
6.4O8
6.408
0.0
O.O
0.0
O.O
O.O
O.O
O.O
0.0
0.0
O.O
0.0
O.O

-------
TEST
NO. : HD-8
SUMMARY REPORT
%
TORO. MAX
RPH
2OOO
2OOO
2OOO
15OO
15OO
15OO
2OOO
2OOO
2OOO
150O
15OO
15OO
FT-LB
89.5
9O.S
89. b
29. O
30. O
3O.O
90.0
90. 0
90.5
29.5
29. O
29.8
TO.
1OO
1O1
100
IOO
103
1O3
IOO
too
101
too
98
1O1
1O872
THR
POSN
MO
OHO
O80
137O
1 16O
87O
MO
O8O
O80
14 2O
125O
9-IO
ENGINE
IGN.
TIMG.
22 .58
22. 8B
23. 2B
44 .OB
44 OB
42 .28
22.58
22. 58
22.08
39. OB
45. OB
4O.5B
INJ.
TIMG.
O.
O.
0.
O.
O
O.
0.
O.
0.
O.
O.
O.
: 380 NAPS-ZM
CORR.
BMP
34
34
34
8
a
8
34
34
34
8
8
8
47
85
44
37
66
66
68
66
82
52
37
61
MEAS.
A/F
8.OO
8 12
817
9.O6
9.79
10.38
8 O3
8. OB
7.95
9. 37
10. 3O
10.97
PHI
EOUIV
RATIO
081
0 80
O 79
071
O G6
O 62
O 81
O.BO
O 81
0.69
O 63
0.59
9O
AIR
LB/HR
208 9
213-1
2125
87 6
98 9
114.6
2112
2134
2112
89 9
99 8
1 16 8
TEST
FUEL
LB/HR
26. 13
26.30
26. 02
9.67
10. 1O
1 1 .04
26.31
26.43
26.55
9.6O
9.69
1O. 65
D/T
BSFC
LB/
BHPHR
O
0.
O
1 .
1 .
1
0.
O.
O
1
1
1 .
758
755
756
156
166
275
759
763
763
127
157
237
: 9-28-82 14: O
BSFC
GAS
EOUIV
LB/
BHPHR
O. 356
O.355
O.355
O.543
O.548
O.599
O.356
O.358
0.358
O.529
O.544
O.581
ENERGY
EFFICIENCY
1OOO*
MBTU/
%
38.9
39.0
39.0
25.5
25. 3
23. 1
38.8
38.6
38.6
26. 1
25.4
23.8
BHPHR
6.55
6.52
6.53
9.99
1O. 08
1 1 .02
6.56
6.59
6.59
9.74
1O.OO
10.69


REPORT


BSHC
1
)
1
4
1O
22
1 .
1 .
1
5.
6
12.
776
621
679
, 4O5
816
8O8
796
769
.799
.229
.744
. 148
D/T:
O9-27-83 13:53 PO6258 1
RAKE SPECIFIC EMISSIONS (G/BHPHR) 	
BSCO
1
2
2
4
5
9
2
2
2
3
5
8
.976
.096
. 154
. 14 1
.993
.047
.212
. 138
. IOO
.358
656
.664
BSC02
453
446
445
616
572
591
448
443
444
626
6O7.
693.
.53
.93
.23
.63
.88
.60
.89
88
.55
62
3O
74
BSNOX
7.787
6.372
6. 136
12.112
4.412
2.O82
7.788
6.324
6.219
4 .457
1 .833
2. 181
BSALDVH BSPART
O.O
0.0
0.0
O.O
O.O
O.O
O.O
O.O
00
O.O
O.O
O.O

-------
TEST NO.:  HO-81O873
                          ENGINE:  380  NAPS-ZM2
TEST D/T:  9-29-82 1O: O
REPORT D/T: O9-27-83  13:57
                                                                                                                             PO6258 1
SUMMARY REPORT
            %                                PHI  -              BSFC
     TORO. MAX THR  IGN.   INJ.  CORR. MEAS. EOU1V   AIR     FUEL  LB/
RPM  FT-LB TQ. POSN TIMG.  TIMG. _BHP_  A/F  RATIO   LB/HR   LB/HR BHPHR  BHPHR  %   BHPHR
                                                                       BSFC     ENERGY
                                                                       GAS    EFFICIENCY
                                                                       EOUIV       1OOO*
                                                                       LB/         MBTU/
2OOO  9I.O 10O   11O 2O.OB   O.  35.O9  7 86  O 82
2OOO  89.8  99  O8O 21.5B   0.  34 63  8.06  0.80
2OOO  9O.2  99  07O 21 .8B   O.  34.SO  7.95  OBI
15OO  29.O 1OO  144O 3O.OB   0.   8.31  8.66  O 75
15OO  28.7  99  12OO 48 OB   0.   8.23  9.97  O 65
15OO  29.5 1O2  930 51.OB   O.   8 46  1O.25  O.63
                                                    211  2   26.87 0.766 O.36O  38.5  6.62
                                                    213  4   26.47 O.764 O.359  38.5  6.61
                                                    212  1   26.69 0.774 0.363  38.1  6.69
                                                    81  8   9.44  1  136 0.534  25.9  9.82
                                                    98  O   9.83  1.194 0.561  24.7  10.32
                                                                                             BRAKE SPECIFIC  EMISSIONS (G/BHPHR) 	
                                                                                                BSCO
                                                                                                       BSC02    BSNOX  BSALDYH BSPART
                                                                                        2.O94
                                                                                        1 .841
                                                                                        1 .837
                                                                                        4.496
                                                                                        5.956
                                   1 .883  449.19  7.O38
                                   2.O42  443.94  6.324
                                   2.O66  450.21  6.387
                                   3.665  616.35  6.1OO
                                   6.145  59O.2O  5.151
                                                    108.7   10.61  1.254  0.589  23.5  tO.B4  11.2OO   8.684   582.28  2.521
                                     O.O
                                     0.0
                                     O.O
                                     O.O
                                     O.O
                                     0.0
                                                                                                                                        CD
                                                                                                                                        I

-------
TEST
NO. : HD-8
SUMMARY REPORT
%
TORQ. MAX
RPM
2000
2OOO
2OOO
15OO
1500
1SOO
2OOO
2OOO
2000
15OO
15OO
15OO
FT-LB
89.5
89. O
90. 0
29.8
29.5
29.4
9O.O
9O.O
89.8
30.5
30.0
28.0
10^
too
99
1O1
10O
99
99
100
too
1OO
too
98
92
1O874
THR
POSN
120
oao
070
145O
134O
990
O90
OBO
oao
1430
137O
1O60
ENGINE
IGN.
TIMG.
21 .58
23. OB
22. OB
44. SB
54 .OB
55. OB
17 .OB
20.58
21 .88
52. OB
47 OB
50. OB
INJ.
TIMG.
O.
O.
0.
O.
O.
O.
O.
O.
O.
O.
O.
O.
: 38O NAPS-ZM2
CORR.
BMP
34
34
34
8
a
8
34
34
34
8
8
8
.39
18
.57
. 58
'IS
.45
.56
.57
49
.78
63
.05
MEAS .
A/F
7.89
8.12
7.93
8.65
9 G5
10.53
7.62
7.83
7.95
8.51
9 47
10. 46
PHI
EOUIV
RATIO
0 82
0 SO
0.82
O. 75
O.67
O 61
O.B5
O.83
O 81
O.76
0 68
O 62
AIR
IB/
2O8
213
21 1
81
89
107
21 1
21 1
213
82
87
103
0

TEST
D/T:
9-29-82 14:
BSFC
GAS
BSFC EOUIV
FUEL LB/ LB/
MR
9
4
.2
a
.9
.a
.2
2
.0
7
.6
3
IB;
26
26
26
9
9
IO
27.
26.
26.
9.
9
9.
'HR BHPHR BHPHR
.49 O
30 O
.62 O
46 1
31 1
24 1
70 O
96 O
79 O
71 1
25 1
88 1 .
.770 O.
.769 O.
.770 O
. 1O3 O
. O98 O
.212 O
.802 O
.780 O.
.777 O
107 0
.072 O.
227 O.
362
361
362
518
516
569
377
366
365
52O
, 5O4
576
0
ENERGV
EFFICIENCY
10OO*
MBTU/
y.
38.2
38.3
38.2
26.7
26.8
24.3
36.7
37.8
37.9
26.6
27.5
24. 0
BHPHR
6
6.
6.
9.
9
10
6
6
g
9
9
10
66
65
66
53
49
48
93
74
71
57
27
61
REPORT




BSHC
2
2
2
8
6
9
2
2
2
8
6
7.
.753
.321
.254
.931
.532
. 161
.827
. 30O
. 30O
.461
.071
955
D/T:
O9-27-83 13:57 PO6258
RAKE SPECIFIC EMISSIONS (G/BHPHR) 	
BSCO
2
2
2
3
3
a
2
2
2
2
3
7.
. 35O
.367
.351
.042
. 8O8
. 186
.328
.269
. 3O5
.835
234
941
BSC02
447
448
448
6O2
578
577
437
425
423.
609.
581 .
6O4.
.81
.51
.39
.59
.29
.20
35
07
25
72
48
71
BSNOX
6.O22
5.87O
5.5OO
1O.454
2.S8O
1 . 1O9
4.783
5.204
5.671
13.727
2.879
O.946
BSALOVH BSPART
0.0
O.O
O.O
O.O
O.O
O.O
0.0
0.0
O.O
O.O
O.O
0.0
to

-------
TEST NO.: HD-8IO877
                         ENGINE: 38O NAPS-ZM2
   TEST  D/T:  10-19-82  13:  O
REPORT D/T: O9-27-B3 13:59
                                                                  PO62S81
SUMMARY REPORT
            %                                PHI
     TORO. MAX THR  IGN.  INJ.  CORP. MEAS. EOUIV  AIR
RPM  FT-LB TO. POSNI T1MG. TIMG. _BHP_  A/F  RAJJ.O  LB/HR

150O  29.5 tOO 142O 44.OB   O.   8.51  8.69  O 74   82.7
150O  29.1  99 1280 48.5B   O.   8.38  9.74  O66   92 1
15OO  3O.O 102  9OO 45.5B   O.   8.64 10.42  O.62  111.4
FUEL
LB/HR
BSFC
LB/
BHPHR
BSFC
GAS
EOUIV
LB/
BHPHR
ENERGV
EFFICIENCY
10OO
UQ Til/
MB I U/
% 	 BHPHR

BSHC
                                   BRAKE  SPECIFIC EMISSIONS (G/BHPHR)  	
                                     BSCO    BSCO2   BSNOX BSALDYH BSPART
 9.52  1.119  0.525  26.3   9.67   3.299
 9.46  1.128  0.53O  26.1   9.75   5.934
10.70  1.237  0.581  23.8  1O.69  19.985
     3.123  599.4O 11.074
     4.TOO  586.13  3.888
     8.4O5  564.77  1.58O
O.O
O.O
O.O
                                                                                                                                      CD
                                                                                                                                      O
                                                                                                                                       I

-------