DIRECT WATER INJECTION COOLING FOR MILITARY ENGINES

DIRECT WATER INJECTION COOLING FOR MILITARY ENGINES
DIRECT WATER INJECTION COOLING FOR MILITARY
ENGINES AND EFFECTS ON THE DIESEL CYCLE
R. B. MELTON, JR., S. J. LESTZ, AND R. D. QUILLIAN, JR.
U.S. Army Fuels and Lubricants Research Laboratory, Southwest Research Institute, San Anlonl.o, Texas
ANn
E. J. RAMBIE
U.S. Army Tank Automotive Command, Detroit, Michigan
A study was conducted on the feasibility" of totally cooling a single-cylinder diesel engine
by direct injection of water into the combustion chanlber. The term "total cooling" can be
taken to mean stabilized cooling at all loads and speeds so as to eliminate need for conventional cooling jackets, cooling fins, or oil spray jets. The engine used was a CLR Direct Injection
Diesel with 42.5 cubic inch displacement and a compression ratio of 16:1. Most of the running
was at 1800 rpm and 92 psi IMEP.
Separate measurements were made of heat rejection to the cylinder tread, liner, and crankease oil to determine more accurately where the cooling effect was being applied. Water
injection was by means of a Bosch pnmp and various pencil-type nozzles installed, adjacent
to the fuel injector in the cylinder head. Port injection and port induction were also briefly
investigated. A five-hole, 90 ~ included angle nozzle was used, as was a three-hole, 30 ~ included
angle unit. For comparison, a nozzle directing one spray obliquely at the cylinder wall was also
tested. Firing pressure was monitored using a piezo-eleetric transducer; both pressure-time and
pressure~volume (indicator) records were obtained. In order to determine timing of both fuel
and water injection, needle lift was monitored using a differential transformer pickup.
The results of this study indicate:
Optimum total engine cooling by direct water injection was accomplished over a wide
range of water injection timings (from 450 to 720 CA degrees after TDC power stroke) at
water/fuel ratios of 2.9 to 3.7 with output power and brake specific fuel consumption improved 5 to 20%, respectively, over that with the standard jacket-cooled CLR engine.
Emissions are affected in an expected manner by the presence of water: NOx is decreased,
sometimes substantially, while the other emissions (HC, CO) tend to increase.
When cooling the exhaust, the condensate becomes an effective scrubber of sulfur oxides.
NO~ was not significantly reduced by scrubbing, but if the condensate is made su~eiently
alkaline (pH :>8), CO2 was unintentionally scrubbed out.
The quail W of the uncondensed exhaust for turbooharglng is attractive. A thenretical gain
of about 17.5% in available exhaust enerKy due to generation of steam was calculated, along
with a temperature decrease of several hundred degrees Fahrenheit.
Water contamination of the lubricating oil varies from negligible to extreme, depending on
injection quantity, timing, and spray pattern. By not directing water at the liner wall, and by
keeping the oil above 212~ one can maintain the oil in a dry condition.
Based on this work, several pertinent recommendations have been made: (1) utilize water
injection fl)r short-duration, very high-output operation which would otherwise be destructive
due to thermal overload ; (2) use water induction cooling in event of loss of conventional liqnid
coolant; (3) utilize exhaust scrnbbing in stationary applications to permit burning of high-sulfur
fuels without producing sulfur oxide emissions; nitrogen oxides could likewise be reduced by the
injection of small amounts of water; and (4) since 2-stroke-cycle engines are an important
category of diesel engines, some work similar to this effort should be done to this engine type;
prospects are good for success, but conditions are apt to be more restrictive.
1389
PRACTICAL SYSTEMS
1390
Introduction
Internal combustion engines employed in
military vehicles are often subjected to very
severe duty cycle under environmental extremes.
Engines with high power-to-weight ratios are
required for operation in the Arctic at --60~
and in the desert up to 125~ One of the major
problems, particularly for armored combat vehicles, is the extremely limited space in which
to install the power package, Cooling airflow path
is seldom ideal, affected by necessary use of
armored grilles, brush screens, guards, and baffles to direct the air for satisfactory cooling.
Military engines operate trader conditions ineluding radiator or heat exchanger damage,
scaling and plugging, low oil and water level and
deterioration which causes reduction of cooling
system performance. Thermal stresses imposed
on engine components and on the engine oil
can result in reduced engine life or curly failure.
Questions addressed in this study of total
cooling by water injection feasibility are as
follows:
(l) Can operation of the diesel engine be
achieved using direct water injection for total
cooling?
(2) What conditions of timing spray and
quantity of water are required?
(3) Can direct water injection be used for
total cooling with 100~v recovery of cooling
water ?
(4) What is the impact of direct water injection on emissions, i.e., NO, NO~, hydrocarbons,
carbon monoxide, and oxides of sulfur?
(5) When the exhaust is cooled to recover the
water, does the condensate function as a scrubber
for some emissions, i.e., sulfur oxides, NO~ or CO?
(6) Does water injection cooling have potential for stationary engine applications?
(7) Can water injection be utilized for short
duratioo trader high-output operation which
would otherwise be destructive due to thermal
overload ?
(8) What is the impact of water injectinn on
lubricant quality, eagute deposits aaxd wear of
engine parts ?
For the foreseeable future, conservation and
enviromnental pressures will cause all researchers
to consider fuel availability, fuel economy and
engine gaseous emissions. In addition to the
reduction in nitrogen oxides (NO=) reported by
others, a-4 the probability existed that oxides of
sulfur (SO~) might also be removed from engine
exhaust during operation with water injection.
If so, this could make fuels with relatively high
sulfur content look attractive, at least, perhaps
for stationary applications. Earlier work2 conducted at our laboratory demonstrated the
feasibility of totally cooling an air-cooled spark
ignition engine by direct water in~eetion. While
the concept of injecting (or inducting) water h~to
an internal combustion engine (SI or CI) is not
new, ~--~it is nevertheless considered that the idea
of assuming the full burden of handling rejected
heat via direct water injectinn and reaping the
advantages is an approach which may contribute
significantly to the potential of internal-combustion engines. In some cases, this is because of
the new types of engines (high-speed and highoutput) that have come along; in other cases
it is because the problems facing the internal
combustion engine designer are somewhat different from those of times past (nitrogen and
sulfur oxides emissions, particularly, and the fuel
situatin~t).
The current work1~ was mldertaken to determine feasibility of total cooling diesel engines by
direct water injection a~td determhm effects on
diesel cycle operation and combustiou. A schematic illustration of a proposed water-injection
condensate recycle concept is shown in Fig. 1.
One of the features of such an engine is the potential of providing for increased water injection
with increased severity of cooling load. A~ illustrated by Fig. 1, liquid coolant is injected directly into the combustion chamber during
compression, expansiov., mid/or immediately
HOT EXHAUST GAS
RADIATOR
CONDENSER
WATER CONDENSATE
~NJECTICN
PUMP
EXHAUST GAB
UNBURNED HYDROCARflON$
WATER CONDENSATE
HEAVY SOL~D5
Fla. 1. Schematic illustration of water-injection
condensate recycle concept.
WATER INJECTION COOLING
following the "effective work" portion of the
combustion cycle. When injected, the coolant
absorbs heat front the combustion chamber,
vaporizes, and is expelled during the exhaust
cycle along with products of combustion. As the
combustion and coolant products flow into the
heat exchanger, the coolant condenses into a
sump and the cooled exhaust gases are expelled
into the air. Condensate from the sump is then
recycled to the injection pump.
Theoretical Considerations
Early work~ successfully applied direct water
injection total cooling to slow-speed, stationary
gas engines. In this work, water was sprayed at
all surfaces exposed to hot gas near top center on
the fim~g stroke. It was important t h a t the
water did not vaporize from the hot combustion
gases, but rather, it was permitted to reach the
hot surfaces from which it would remove heat
by formation of steam. The injected water required was in close agreement with the normal
engine heat rejection divided by the heat of
vaporization water.
Utilizing the concept of vaporization at solid
surfaces as the basis for cooling effectiveness, it
is possible to examine some of the implications
of t h a t "mode]." In the absence of a n y known
cogent model built to explain the direct effects
of water injection on the combustion gases, the
simplest analysis, based on the conservation of
energy, can be applied. For example, assume a
diesel engine has a thermal efficiency of 33.3v/v,
and is operating with a fuel having a lower heating value (Qs) of 18 000 Btu/Ib. Distribution of
supplied heat, Qs, is normally divided equally
between the exhaust gas (6000 Btu/lb), the
jacket cooling load (6000 Btu/lb, including
lubricating oil), and shaft work (6000 Btu/lb).
Then theoretical water required (Wj ) to pick
up the jacket heat load, Q j , at total cooling
would be
W z = Q j / & H = 0.333 (18 000)/1100
= 5.5 lb water/lb fuel
where &H = 1100 Btu/lb = average heat of vaporization for water between 200~ and 400~
A possible contribution of the current work
to the theoretical understanding of water injeetion cooling is recognition of the effect of
pressure in the combustion chamber on the
cooling effect of water, n During a significant
part of the diesel cycle (compressinn and power
strokes), one cannot expect boiling film cooling
1391
to occur because ambieut pressure exceeds saturation pressure. If typical metal temperature is,
say, 300~ then no boiling will occur when the
pressure exceeds 67 psia. Above this pressure,
water will simply wet the walls, warm up, and
await the decline of cylinder pressure to the
point that boiling carl occur. Thus, vaporization cooling can occur only during a portion of
tbe compression stroke and the latter part of
the power stroke. The only benefit obtained
from injection during the early power stroke
comes from the fact t h a t the piston is close to
the head during t h a t time a n d the injection
targeting may therefore be easier. This part of
the cycle (power stroke) has high pressure; high
injection pressure is required in consequence,
and there will be a tendency for excessively
atomized water to vaporize in the gas phase,
rather than at the walls. As a result, excessive
amounts of cooling water will be needed, while
the charge cooling will cause increased ignition
delay or substantial power loss just after cmnbustion. None of these effects is desirable.
A question may be asked whether additional
work is produced by the generation of steam
during the power stroke. It appears from analysis
of tliis situation that the pressure gain from steam
formation is about one-sixth of tile pressure loss
due to charge cooling; thus, a n y vaporization
in the gas phase during the power stroke constitutes a pure loss.
The other way to get work is to have steam
formed by vaporization of water at the walls;
this steam is "free" and is potentially a source
of increased power. A calculation indicates that
one can expect a 3-psi increase in mean effective
pressure (assuming about 75 psi IMEP) if it
were possible to get steam all formed halfway
through the power stroke. Recalling the difficulty previously explained where belling is suppressed by firing pressure, it is not probable t h a t
more tha~l half of the expaiLsion stroke can be
utilized. However, it will be sho~al that when
fuel injection timing is advanced to compensate
for increased ignition delay (during total-cooling
water injection operation), the combined diesel
cycle effects actually can produce a significant
increase in power output.
Apparatus
The engine* used in this work was the standard
CLR (with the standard direct-injection diesel
conversion package) coupled to an eddy-currenttype dynamometer. The fuel system used a
* Bore and stroke=3.80 in.X3.75 in. =42.5 CID.
1392
PRACTICAL SYSTEMS
showing the location of fuel injecter and water
injector in relation to the piston cup. The exhaust
system was modified as required to facilitate
recovery, recycle, scrubbing, or analysis of the
exhaust water, whatever the case being investigated.
FUEL
INJECTION
.
NOZZLE
I
Calorimetry System
WAFER
INJECTION
NOZZLE
, -S S P R A Y S
OFFSET ANGLE
INCLUDED ~
o i
/ ' ~ w ~A~ E t
16(] ~
.70~
~
\
TOP
OF
PISTON
Fro. 2. Engine head and wager injector location.
four-hole fuel injector (4X0.010-1n. hole openings, equally spaced, 150 ~ included angle, with
20 ~ offset angle), having a 3000-psi crack pressure. A Bosch APE-1B fuel injection pump was
used with a 6-ram delivery valve and 10-ram
pump plunger. Fuel was supplied from an overhead burette with delivery rates taken in minutes
for 100 ml of fuel consumed and reported in
lb/min. The water-injection system used several
slender, "pencil"-type (0.375 in. diam) nozzles
manufaetm~ed to desired specifications. A Bosch
refit pump (APE-IB) identical to the fuel injection pump was used to drive tim water injector.
The geometry of the water injectors used in this
work is show~ as follows:
When this program was still in the plam~ing
and setup stages, it was expected that total cooling might be difficult to accomplish and that
some means of quantitatively determining the
ex~tent of "partial" cooling would contribute
guidance to the difficult optimization process
that was anticipated to be necessary to achieve
total cooling. The logical indicator of partial
cooling effectiveness is a calorimetric deterruination of absolute heat rejection. It was decided to obtain separate heat flux measurements
for the head coolant, liner coolant, and lubricating
oil. Separate systems were developed for each
of these based on appropriate measurements of
fluid flow rates, temperature changcs, etc. so
that rigorous heat balances could be made. Observed differences in heat-rejection measuroments made during engine operation were found
to be reasonable and repeatable. Diethylene
glycol was used for coolant temperatures up to
230~ and a vegetable oil was used at 325~
and above. Engine metal temperatures were
monitored at the top of the cylinder liner and
in the fire deck of the cylinder head.
Pressure and Timing Measurements
Pressure-time, pressure-volume, and fuel and
water injection timing data were taken with
conventional electro-mechanical devices and
monitored on an oscilloscope.
Exhaust Emissions Measurements
Designation
H40
L150, H150
"A"
Hole
size,
in.
0. 012
0.010,
O. 008
0.051
Number Offset Included
of
angle, angle,
holes
deg
deg
5
3, 4
1
25
30
90
30
O
0
The numbers in the desi~lation represent the
cra~lk angle from TDC at which the spray is
expected to be directed into the piston cup.
Figure 2 is a section through the cylinder head
In all cases, bag sampling was used for basic
emissions measurements. The instruments used
oil this bag sample were as follows:
NO
NOt
CO
COs
02
HC
Smoke
Chemiluminescence analyzer
Chemiluminescence analyzer
NDIR
NDIR
Polarigraphic-type analyzer
Flame type ionization detector
apparatus
Bosch paper filter type with photooptical reading unit
Since sulfur content is an important factor
WATER INJECTION COOLING
in the fuel industry and since sulfate showed up
in the water recovered from the e.xhaust, it
was decided to conduct exhaust scrubbing tests
to remove oxides of sulfur (SO,). Wet-chemical
analysis was used since no direct reading instrument was available.
For a more thorough description of test apparatus, the reader is referred to Re/. 10.
,200
1393
H4ONOZZL[ 325 F COOLANT
o
Experimental Results and Discussion
C,~oling Via Direct Water ln:jection
Utilizing the apparatus as previously described,
the engine was operated under the following
fixed conditions, /mless specified otherwise:
Spend: 1800 rpm
Fuel rate: Set to maintain constant brake load
of 19.2 lb ~dth no water injection (5.2 Bhp at
1800 rpm) (Approx 0.056 lb/min). Fuel rate with
water injection was also equal to 0.056 lb/min.
Only fuel injection timing was varied and not
fuel delivery rate. Timing advance l~equired to
maintain TDC as the end of the iguition delay
period ranged from 6 to 18 CA degrees over that
of the standard jacket-cooled engine.
Water injection
Rate: 0 to 0.40 lb/min
Timing: Varied hy 90 ~ CA increments
Duration: Varied from 30 ~ to 90 ~ CA degrees
Spray characteristics: Nozzles were chosen to
provide spray angles which would be targeted
for full spray entering the pistion cup at 40 ~ to
150 ~ from TDC utilizing cone angles of 30 ~ and
90 ~ Iu addition, one single spray nozzle of 0 ~
cone angle was used.
Since three different heatflow measurements
were obtained and since the cooling effect was
not equally divided among head, liner, and oil
systems, their sum was used to define total
cooling. When net heat rejection from the engine
equaled zero total absolute cooling was achieved.
In practice, net heat flow had to be computed
after-the-fact and the total-eoolfilg condition
estimated by linear hlterpo/atiou because all
tests were conducted using predetermined water
injection rates which nfight have been slightly
more or less than that required for total cooling.
Space does imt permit discussior~ of the individual heat rejection data acquired for the cylinder
head, cylinder liner, and sump. Figure 3 shows
total engine heat rejection for the five-hole nozzle
versus water/fuel ratio for the full range of
injection timings. Total cooling water/fuel ratio
is that ratio at which the net heat rejection ia
zero. Interestingly, the range in water/fuel ratio
100
I
I
I
I
~ qOeOi
2
3
4
5~4~6
WATER/FUEL IW/F) RATIO. LB'LB
7
FZG. 3. Total Btu rejected versus water/fuel
raticr--Numbers on curves are the end of waterinjection timing, CA degrees from TDC power
stroke.
is from about three to somewhat over five.
Maxinmm water/fuel ratio of 5 is required for
total cooling when injecting during the first 90
deg of the power stroke. A minimum water/foe[
ratio of 2.9 results for injection during the last
half of the intake stroke. In all probability,
the former behavior results from unnecessary
cooling of dense, hot gas (there is a marked power
loss), while the latter is associated with spraying
water directly onto the liner (which will be seen
to cause a great deal of water to enter the
crankcase).
Figure 4 shows the interpolated water/fuel
ratios on a weight basis for total cooling plotted
versus the approximate end of the water injection event in degrees ATC (power stroke) for
the 1[40 and L150 water injection nozzles.
Water injection tinting has an important effect
on required water/fuel ratio for total cooling.
Total cooling is possible at any injection timing
tried here. Since injection duration is a significant part of 90 deg, it m a y be said that no
part of the cycle has been skipped, and further,
that there is a slight sensitivity to water injection spray geometry.
Ush~g the total cooling water/fuel ratio, Fig.
5 is obtained showing brake load and specific
fuel consumption plotted against water injectiou timing. Note that the power output was
obtained at constant fuel rate (0.056 lb/miu),
and therefore increases in engine brake power
result in decreases in brake specific fuel consumption. Notice that the instant water is injected into the hot, dense gas (middle of power
stroke) there is a drastic decrease in power and
fuel economy. This demonstrates that injection
PRACTICAL SYSTEMS
1394
for injection during the exhaust stroke to 20.8%
for injection at the end of the compression stroke
are realized, compared with the standard jacketcooled engine. This shows that evaporative
cooling of the metal surfaces is a more efficient
process and produces beneficial effects in the
cycle. Also note t h a t the fue[ consumption improves from 0.65 to 0.54 l b / B h p / h r for the best
water injection condition.
Figure 6 shows how water collected from the
crankcase "blowby" (using a copper coil condenser) varies with water injection timing at the
total-cooling water/fuel ratio. Since oil temperature exceeds 212~ water bypassing the rings
will not accumulate but will be evaporated and
expelled. As can be seen, the quantity of water
T
r
Ig0 BTU/MIN PER 0.2 LE/MIN = ~o0 STO/LB
- - LISO
NOZZLE
325"F COOLANT
ExHAusT
INrAKE
POWER
,
0
E
=i~,
I
=i~,
-
COMP'N
I=i-
I
100
200
300
4OO
EO0
co0
END OF WATER INJECTION TIMING, ATC
J~
_1_
_1
1
-POWER T , - EXHAUST'- INTAKE :'=COMP'N
'l
7O0
FIG. 4. Water/fuel ratio for total cooling versus
water iDjectioa timing.
into the hot gases causes charge cooling and has
a detrimental effect on the cycle. However, if
water injection is confined to the balance of the
cycle--exhaust, intake, or compression strokes-power and economy increases ranging from 5.2~o
o
YI,
IO0
, II
2O0
3O0
4O0
,1
50O
,
6O0
END OF WATER iNJECTION TIMING, ~
H40 NOZZLE, 325~F
=_, 20
.--._.
17
--~
0.65
,s
0.6o
13
0.50
12
0.45
11
POWER
,or
O
,'i~,
EXHAUST
,~F,
7OO72O
ArC
FEO. 6. Water rejection to sump at total cooling
versus water injection timing.
NTAK
,=F,
too
2OO
300 ~
500
8OO
END OF WATER INJECTFONTIMING, ATC
:'o,o
70O 720'
Fro. 5. Brake load a~ad BSFC at total cooling (interpolation of Fig. 3) versus water injection timing.
going this way m n ~ s from 0.5 cc/min during
the power stroke to 9.5 ce/min during the compression stroke. Probably the water cools the
parts of the liner where it hits, and the rings
pump this into the crankcase during the next
stroke. Injection during the exhaust stroke has
less tendency to flood the crankcase than intakecompression injection. The 30 ~ included angle
nozzle (L150) produced essentially the s~me results (data not shown), which is somewhat remarkable, considering how much the two nozzles
differ.
Cooling via port injection and port induction
were also studied on a limited scale. Figure 4 and
Table I show the water/fuel ratio of 3.36 (required for total cooling with port injection) superimposed over the direct inieetion water/fuel
ratio requireme.ts. In this instance, water was
injected into the inlet port during the last of
the compression stroke, and thus was all on the
WATER INJECTION COOLING
port walls by the time of the next intake stroke.
It appears that port injection cooling is about
equivalent to direct injection cooling but simpler
to achieve for practical purposes. Of equal
significance, there was a 10~ increase in brake
power and fuel economy, and less than 1 cc/min
of water fotmd its way to the engine sump.
The experiment with port induction was conducted to deternfine the amount of water the
engine could tolerate. At a water/fuel ratio equal
to 3.5, there was a 10~0 increase in power a~ld
fuel economy. Further increases in water input
caused very rapid rates of pressure rise to occur
at water/fuel ratio equal to 5, with power down
1 5 ~ at water/fuel ratio equal to 6, and a 50%
power loss at water/fuel ratio equal to 9. Neither
calorimetry nor emissions were taken during this
test.
Following tests to determine requirements for
total cooling by water injection which were conducted with the calorimetry system, it was
desired to demonstrate total cooling with direct
water injection without coolant flow in the head
and block. This was done and the results are
shovnl in Fig. 7. During these tests, the water/fuel
ratio was varied and the effects were determined.
The data in Fig. 7 suggest that lower water/fuel
ratios are possible as critical temperatures are
being raised.
Table 11 presents a summary of the totalcooling picture with emphasis placed on the more
TABLE I
Port injection test summarized
Conditions:
1800 rpm
20 lb load without water
Water flow rate: 0.188 lb/Min
Fuel flow rate: 0.056 lb/Min
Water/fuel ratio: 3.36 lb/lb
Water injection during last of compression, into
intake port
Circulating coolant and oil temperature: 325~
Resulls :
Load: 22.0 lb with water, up 10%
Head heat flow: +6 Btu/Min
Liner heat flow: --3 Btu/Min
Sump heat flow: 0
Net heat flow: +3 Btu/Min, undercooled
Deck temperature: 456~
Liner temperature: 349~
Exhaust temperature: 570~
Sump water flow: Less than 1 ml/min
1395
~
50(
O=s1509
= H150NOZZLE,iNJ DURING
t g ~ tg~Tg0*OFF~R STROKE
COOLINGLAST
=
D E C K ~ X= TOTAL
L150STROKE
NOZZLE-W/F
= 3.9go
AT
PWR
COOLANT"
325~
4~
/
....
i ...p___
/
,
\
p,,..\,
_
* H E N USED
_
~"
_
--~_
I
WATEWFUEL RATIO.LS/L8
FIG. 7. Estimated equilibrium temperatures versus water/fuel ratio with head and liner coolant
drained.
favorable or optimal times in the cycle for water
injection. From this it is seen that optimum
conditions range from 450 to 720 CA deg after
TDC power stroke at wator/fuel ratios of 2.9 to
3.7 with output power and fuel economy improved 5 to 20 percent, respectively, over that
with the standard jacket-cooled CLR engine.
Effects on the Diesel Cycle
The effects of direct water injection on the
diesel work cycle were investigated at waterand fueMnjection timings believed likely to
exhibit significant effects. For lack of space, the
following discussion of cycle data concentrates
on the most optimal time in the cycle for water
injection--during late compressinn stroke--and
summarizes the effects observed during water
injection at two different timings during the
power stroke. These conditions represent the
extremes of strong combustion influence, i.e.,
charge cooling and 11o charge cooling effect:
(1) " T D C " - - T h i s puts water injection in the
range of 85 ~ to 30 ~ BTC and therefore just
prior to fuel injection. General results are a
substantial increase in ignition delay, strongly
1396
PRACTICAL SYSTEMS
TABLE II
Optimum engine selection chart at total internal cooling conditions
End of water injection
timing, ~ ATC
power stroke
Percent increase
in brake power
and fuel economy
Water fuel
ratio
required
5.2
12.5
13.5
20.8
10
2.9
3.1
3.3
3.7
3.36
Direct injection
450
540
630
720 (TDC)
Port injection 720 (TDC)
effective cooling, and a power and BSFC gain,
with this engine (Fig. 8).
(2) "90-deg A T C " - - T h i s involves injecting
water into dense, hot gas over the range of 5 ~
to 80 ~ ATC and inevitably causes a substantial
power loss. Much water is required for total
cooling, and the increase in ignition delay is
relatively small.
(3) "180-deg A T C " - - T h i s timing injects water
during approximately the last 90 deg of the
power stroke, where its presence is not felt so
strongly. The total-coolhlg requh-ement is mod-
I~
ps~
Sump water
condensate
cc/min
1.6
7.2
9.5
5.5
<1.0
erate, power loss is nil, and ignition delay increase is minor.
Figure 8 shows the situation relative to the
no-water case for water injection late in the
compression stroke, just preceding fuel injection.
Show~l in that figure are the whole i~tdicator
diagram, the magnified bottom of same, and
the pressure-time situation (plus injection timings) in the vicinity of power stroke TI)C. The
flfll indicator diagram shows lower pressures to
occur late in the compression stroke and early
500
R
BEFORE TDC I - - - - - - )
COMPARED WITH CONDITION
OF NO-WATER INJECTION ( - - . - - I
H40 NOZZLE. 325~ COOLANT
LOAD 20 LR WITHOUT WATER INJ,
LOAD 21 .S WITH WATER INJ,
WATER/FUEL RATIO-3.R
"~
DELAY
~
EXPANDED BO'FrOM
~
! , .Eo :
RECORD
-"
0.o01 5EC
COMPENSATED}
SID LESTZ, FIGURE B, REDUCE
TO 45% OF ORIGINAL
Fro. 8. Cycle data-wate~' injection immediately prior to TDC.
WATER INJECTION COOLING
1397
TABLE I I I
Engine emissions with various water-injection timings
Date
(1973)
21 Nov.
26 Dec.
21 Nov.
26 Dee.
21 Nov.
26 Dec.
21 Nov.
26 Dec.
NO,
ppm
NO~
ppm
HC,
ppm
CO,
%
450
590
75
238
115
560
585
115
257
160
403
610
279
700
530
753
194
710
182
0.80
0.60
0.41
0.58
1 .Ol
0.30
1.26
0.54
377
200
412
240
428
Bosch
smoke
No.
CO2,
%
-6.2
-2.2
-5.8
-6.4
5.2
6.2
6.9
8.5
9.4
7.1
8.7
7.6
8.9
9.6
0.35
Timing
Timing
Timing
Water Approxinjection imate
0~,
nozzle water/fuel
%
type
ratio
9.0
8.3
12.0
9.35
11 .O
9.15
7.0
8.1
360~
450~
540~
None
None
A
H150
A
H150
A
HI50
H150
H150
H150
0
0
4
3
4
3
4
3
3
3
3
Comments
Three runs average
One run only
Timing last part
Compression stroke
Timing first part,
Power stroke
Timing l~st part,
Power stroke
(Smoke data only)
(Smoke data only)
(Smoke data only)
1800 rpm, 20-1b load, fuel rate about 0.05 ib/min temperature 325~ for oil and coolant (vegetable oil).
in the expansion stroke with some slight gain in
peak pressure late in the expansion stroke. The
indicator and pressure-time (P-t) diagram show
a net increasc in cycle work as a result of the
following:
(1) Reduced compression work due to ongoing cooling of the working fluid. This cooling,
however, also reduces early expansion work by
a similar amount.
(2) Increase in end-of-cycle (late power stroke)
work due to flash evaporation of water from the
hot metal smffaees. Evaporation is suppressed
until cycle pressure falls below the water saturation pressure corresponding to the cylinder metal
temperature.
(3) The P-t record shows conditions near-TDC
which are not evident from the indicator diagram.
We see the reduced compression pressure noted
above. Also, the increased firing pressure just
after TDC augments cycle work. Iguitinn delay
was increased by water injection, but was compensated for by advancing fuel injection timing.
The ignition delay is attributed to reduced compression temperature as well as possible chemical effects. It is considered that with this engine,
the increased ignition delay with water injection
improves fuel mixing and combustion, accounting
for a substantial part of the observed power
increase.
Volumetric E~ciency
The effect of water injection (using the 5-hole
H40, 3-hole L150, and shlgle hole "A" nozzles)
and port induction on volumet~bc efficiency was
studied, in general, it was observed that water
input caused a slight but sustained increase or
decrease (2-3%) in volumetric efficiency. Increases occur due to charge and en~ue metal
cooling, while decreases are due to the formation
of steam on the intake stroke. However, in this
study significant transient effects were observed
in that commencement of water input ofteu
produced a much greater reduction in volumetric
efficiency (19-12%) and cessation of water input
typically produced a temporary boost (a few
percent) in volumetric efficiency.
The Engine Exhaust
Emissions. Table I I I shows the exhaust emissions produced at standard speed and load by
the CLR ealghm without or with water injection
for three water-injection timings and two wateriniection nozzles. Emissions data were not taken
at any other times in the cycle. The following
comments can be made about the data:
(1) The basic engine exhaust contains a great
deal of oxygen; the volumetric efficiency is low
and so is the utilization of what air is taken in.
Smoke is high, as is CO.
(2) Injection of water during the last part
of the compression stroke seems to cut NO= substantially, quite in keeping with expectations.
In one case shown, smoke declined sharply while
CO was not changed. Why oxygen should increase is not known.
1398
PRACTICAL SYSTEMS
(3) Injection of water during the first part
of the power stroke (into the hot high-pressure
gas) produces a fair effect on NO=. Whether this
is due to water carry-over into the combustion
phase is not known for certain, of course. Hydrocarbons went up for one water injector and
down for another, while smoke was unaffected.
The decrease in CO is unexplained; the increase
is more expected.
(4) Injection of water late in the power stroke
reduces NO= while the effect on hydrocarbons
parallels that of the earlier timing. Smoke was
tmaffected.
Direct water injection has the greatest effect
on combustion when injected during the late
compression stroke. This is partially due to the
increase in ignition delay caused by charge
cooling, allowing greater mixing prior to ignition. Also, the resulting combustion is influenced
by the large amount of water present. For the
other two water injection timiugs, the water is
introduced after the combustion event, and
therefore, does not affect the course of combustion. However, carry-over effects are likely to
affect emissions. It appears both experimentally
a n d theoretically t h a t parts remain wet during
the entire cycle, and that water vapor can get
into the intake charge and chemically influence
combustion rather than affecting it by cooling
the compressed charge and causing increased
delay as described above.
It is worth noting that the changes in smoke
and CO were independent, instead of being
closely linked as normally expected. And finally,
engine emissions values were reasonable, allowing
for poor volumetric efficiency and air utilization
of the CLR engine.
Engine emissions as affected by injection of
aqueous solutions. Several aqueous solutions were
directly injected into the engine combustion
chamber to compare their effects on exhaust
emissions with the injection of plain water; results
are in Ref. 10.
Exhaust scrubbing. An extensive study of exhaust gas scrubbing (after the exhaust port) was
conducted, but space permits reporting only the
highlights.
(1) None of the solutions attempted scrubbed
NO and NO= significantly. These solutions ineluded sodium bicarbonate solution, sodium
carbonate solution, strong caustic soda solution,
weak peroxide solution, and neutral and acid
solutions. The problem is t h a t NO must be oxidized to N(:h before it can hydrolyze, and even
that process is inefficient.
(2) Scrubbing of exhaust sulfur seems to be
very effective regardless of acidity or alkalinity
of the scrubbing solution, although only about
i
i
i
t
i
i
100~RECOVERY
__ 180
"
~
ZERO MAKEUP
145
S
RETICAL~~~oF/ARel
~ 03
>:120 -ITHEO
r
COMBUSTION
-O~e'--%%~5~
-TOTAL
WATERWATERT -190
\',.,
E
i
i
110
J
i
i
i
120
130
140 150
EXHAUST TEMPERATURE. ~
9 '
180
170
F[o. 9. Exhaust water recovery rate versus exhaust
temperature (single experiment).
45% of theoretical sulfur has been found in the
exhaust by wet chemical methods. It is possible
the remainder of the sulfur was combined with
particulate carbon in the exhaust smoke, and
being in the form of inert carbon-sulfur complexes, CS~ (Refs. 12, 13, 14, and 15), it could not
be scrubbed out with the gaseous sulfur compounds. Indications are t h a t sulfur oxides will
indeed be scrubbed out in ahuost a n y wet system,
although our work did not result in a complete
sulfur balance.
Other exhaust studies. Exhaust water recovery
and the aspects of turbocharging with direct
water injection were also examined. Again, space
does not permit discussion of these, and the
reader is referred to Ref. 10 for these details,
which conclude that:
(1) Water can be recovered quantitatively
from the exhaust in agreement with hunfidity
table predictions; the exhaust must be cooled
nearly to 100~ at atmospheric pressure for 100%
recovery. This is shown graphically and compared with the theoretical case in Fig. 9.
(2) Turboeharged engine calculations show a
theoretical gain of about 17.5% in available
exhaust energy due to generation of steam; exhaust temperature will be reduced several hundred degrees at the same time.
Conclusions
1. Direct water injection cooling of diesel
engines can be accomplished with increased
power and better BSFC. Optimum total engine
cooling by direct water injection was accomplished over a wide range of water injection
timings (from 450 to 720 CA deg after T D C
power stroke) at water/fuel ratios of 2.9 to 3.7
with output power a n d brake specific fuel consumption improved 5 to 2 0 0 , respectively, over
that with the standard jacket-cooled CLR engine.
W A T E R I N J E C T I O N COOLING
2. Total cooling of military diesel engines for
ground mobile application is not feasible due to
the required condensing temperature for full
water recovery (approximately 100~ at atmospheric pressure). T h e totally cooled water
injection engine is impractical as a military
engine t h a t must operate on a self-sustaining
basis.
3. Emissions are affected in an expected manner by the presence of injected water: NO~ is
decreased, while HC and CO tend to increa~.
4. When the exhaust is cooled to recover the
water, the condensate functions as a scrubber
for some emissions. Sulfur oxides are effectively
scrubbed, while NO, and CO are not removed
from thc exhaust.
5. The possibility of control of sulfur oxides
m a y justify use of a wet exhaust system for
stationary applicatious to permit high-sulfur
fuels in the future.
6. Short-duration direct-water injection (a
more traditional role for water injection) m a y
have t•
use for spurt power or overload
operation of military vehicles without water
recovery.
7. Long-range effects of direct water injection
on engine lubricant quality or hardware durability
were not determined, but results show t h a t
advcme effects arc minimized if excessivc water
application to the cylinder wall is avoided. I t is
also important to keep the oil above 212~ to
maintain the engine oil in a dry condition.
2.
3.
4.
5.
6.
7.
8.
9.
Acknowledgments
10.
This work wa~ sponsored by the U.S. Army,
Contract D.a~AD05-72-C-0053, under technical
management of E. J. Ramble, USA TACOM.
Contract monitor was C. F. Schwarz, USA MERDC.
The authors also thank E. R. Lyons who provided
the laboratory assistance, and other members of
the U.S. Army Fuels and Lubricants Research
Laboratory who provided many helpful sugg~tions.
11.
REFERENCES
1. NICHOLLS~J. E., EL-]~ESSIR[, I. A., AND NEWHALL, II. K. : Inlet Manifold Water Injection
12.
13.
14.
15.
1399
for Control of Nitrogen Oxides--Theory and
Experiment, Paper 690018 presented at SAE
International Automotive Engineering Congress
and Exposition, Detroit, Mich., January 1969.
WEAT,~:R~'ORD JR., W. D. AND Q~mLTAN JR.,
R. D.: Total Cooling of Piston Engines by
Direct Water Iniection, SAE Paper 700880,
November 1970.
L~STZ, S. S., MEYER, W. E., ANn C o L o ~ , C.
M.: Emis~ions from a Direct-Cylinder WaterInjected Spark Ignition Engine, SAE Paper
720113, January 1972.
LnSTZ, S. SI AND Mn~nn, W. E.: The Effect of
Direct-Cylinder Water Injection on Nitric Oxide
Emission from an S.I. Engine, Proc. 14th
FISITA Congr., London, England, June 1972,
pp. 2/34-2/41.
HOPKI~SON, B.: A New Method of Cooling Gas
Engines, British I.M.E. Proc., July 1913, pp.
679-695.
R[CA~DO, H. R.: The High-Speed Internal Combustion Engine, pp. 37, 314, 381, Blaekie and
Son, Ltd., 1955.
TAYLOR,C. F. : The Internal Combustion Engine
ia Theory ar~ Practice, p. 440, John Wiley
and Sons, 1960.
O~:RT, E. F.: Detonation and Internal Coolants, SAE Quarterly Trans., Janlmry 1948, pp.
52-59.
VALD~.~NIS, E. ANn WULFNORST, D. E.: The
Effects of Emulsified Fuels and Water Induction on Diesel Combustion. SAE Paper 700736,
Vehicle En~issions, III, PT-14, 1970.
M~L~ou JR., R. B., LESTZ, S. J., A~D QUILIaAN
JR., R. D.: Direct Water Injection Cooling for
Military Engines, U.S. Army Contract
DAAD05-72-C-0053, USAFLRL Report No.
32, DDC No. AD 780175, March 1974.
KnENA~, J. H. ANn KsY~s, F. G.: Thermodynamic Properties of Steam, 1st Ed., J. Wiley &
Sons, 1963.
DALZ, J. M.: U.S. Patent 3,304,158, February
1967.
M~ULLO et al.: U.S. Patent 2,689,783, September 1954.
TULL~R et al.: U.S. Patent 3,042,503, July 1902.
GonMLEZ: U.S. Patent 3,071,445, January 1963.
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