U.S. patent application number 10/240405 was filed with the patent office on 2003-09-25 for gas compressor.
Invention is credited to Coney, Michael Willoughby Essex, Huxley, Richard Aidan, Morgan, Robert Edward.
Application Number | 20030180155 10/240405 |
Document ID | / |
Family ID | 9888910 |
Filed Date | 2003-09-25 |
United States Patent
Application |
20030180155 |
Kind Code |
A1 |
Coney, Michael Willoughby Essex ;
et al. |
September 25, 2003 |
Gas compressor
Abstract
A reciprocating gas compressor in which a piston (3)
reciprocates within a compression chamber (1) in order to compress
gas. Water is injected through nozzle (6, 7) into the compression
chamber to cool the gas during compression. A source of pressurised
liquid (13) is arranged to accelerate liquid through the nozzle (6,
7). At least one duct (14, 17, 19) connects the pressurised source
(13) to the nozzles (6, 7). The dimensions of the duct (14, 17, 19)
are sized to define the inertia of the liquid to control the
acceleration of the mass flow through the nozzles during
compression such that the cooling capacity of the liquid in the
compression chamber increases as the pressure therein approaches
its final value.
Inventors: |
Coney, Michael Willoughby
Essex; (Swindon, GB) ; Huxley, Richard Aidan;
(Swindon, GB) ; Morgan, Robert Edward; (Lancing,
GB) |
Correspondence
Address: |
BIRCH STEWART KOLASCH & BIRCH
PO BOX 747
FALLS CHURCH
VA
22040-0747
US
|
Family ID: |
9888910 |
Appl. No.: |
10/240405 |
Filed: |
May 27, 2003 |
PCT Filed: |
March 30, 2001 |
PCT NO: |
PCT/GB01/01457 |
Current U.S.
Class: |
417/228 |
Current CPC
Class: |
F04B 39/062
20130101 |
Class at
Publication: |
417/228 |
International
Class: |
F04B 039/04 |
Foreign Application Data
Date |
Code |
Application Number |
Mar 31, 2000 |
GB |
0007927.7 |
Claims
1. A gas compressor comprising a compression chamber to contain gas
to be compressed, a compression piston to compress the gas by
movement of the compression piston in the compression chamber,
valve means for allowing compressed gas to be drawn from the
compression chamber, an atomiser for spraying liquid into the
compression chamber to absorb heat from the gas during compression,
a pressurised source of liquid and a duct arranged to feed liquid
from the pressurised source to the atomiser, wherein the source is
arranged to accelerate the liquid through the atomiser into the
compression chamber and the dimensions of the duct are sized to
define an inertia of the liquid therein to control the rate of
acceleration of the mass flow through the atomiser during
compression such that the flow rate of liquid into the compression
chamber is substantially reduced when the pressure difference
between the source and the compression chamber is high and
substantially enhanced when the pressure difference between the
source and the compression chamber is low.
2. A gas compressor according to claim 1, wherein a valve is
provided to control the timing of the flow through the
atomiser.
3. A gas compressor according to claim 1 or 2, wherein the
dimensions of the duct are sized to define the inertia such that
the mass flow increases through the nozzle in a period of at least
30.degree., preferably at least 45.degree. and more preferably at
least 60.degree. of crank angle of a nominal crank shaft driving
the piston.
4. A gas compressor according to any one of claims 1 to 3, wherein
the pressurised source comprises a reservoir of liquid pressurised
by compressed gas.
5. A gas compressor according to claim 4, further comprising means
arranged to feed compressed gas from the compression chamber to
pressurise the liquid in the reservoir.
6. A gas compressor according to claim 4 or claim 5, wherein the
reservoir is an accumulator, and further comprising a separator
upstream of the accumulator which receives gas and liquid from the
compression chamber, means to feed the liquid in the separator to
the accumulator, and a cooler to cool the liquid being fed from the
separator to the accumulator.
7. A gas compressor according to any preceding claim comprising a
further duct arranged to feed liquid from the reservoir to the
atomiser.
8. A gas compressor according to claim 6, wherein the dimensions of
the further duct are sized to define an inertia of liquid therein
different to the inertia defined by the dimensions of the duct.
9. A gas compressor according to any preceding claim, comprising a
further atomiser for spraying liquid into the compression chamber,
the flow through the further atomiser being controlled by a further
valve.
10. A gas compressor according to claim 9, wherein a branch duct
connects the duct with the further duct.
11. A gas compressor according to claim 10, wherein a third valve
is provided in one of the ducts upstream of the location where the
branch duct joins the duct, while the valve and further valve are
downstream of the branch duct.
12. A method of controlling the temperature of gas in a gas
compressor during compression, the gas compressor comprising a
compression chamber to contain gas to be compressed, a compression
piston to compress the gas by movement of the compression piston in
the compression chamber, a pressurised source of liquid and a duct
arranged to feed liquid from the pressurised source to the
compression chamber, the method comprising the steps of spraying
liquid into the compression chamber by allowing the source to
accelerate liquid through the duct and controlling the rate of
acceleration of the mass flow into the compression chamber during
compression by sizing the duct to define an inertia of liquid
therein such that the flow rate of liquid into the compression
chamber is substantially reduced when the pressure difference
between the source and the compression chamber is high and
substantially enhanced when the pressure difference between the
source and the compression chamber is low.
13. A gas compressor comprising a compression chamber to contain
gas to be compressed, a compression piston to compress the gas by
movement of the compression piston in the compression chamber,
valve means for allowing compressed gas to be drawn from the
compression chamber, an atomiser for spraying liquid into the
compression chamber to absorb heat from the gas during compression,
a pressurised source of liquid and a duct arranged to feed liquid
from the pressurised source to the atomiser, wherein the source is
arranged to accelerate the liquid through the atomiser into the
compression chamber and the dimensions of the duct are sized to
define an inertia of the liquid therein to control the rate of
acceleration of the mass flow through the atomiser during
compression such that liquid can be injected into the compression
chamber when the pressure in the compression chamber is greater
than the pressure of the source.
14. A gas compressor according to claim 13, wherein the inertia is
such that the period during which liquid is injected while the
pressure in the compression chamber is greater than the pressure of
the source is at least 5.degree., preferably at least 10.degree.
and more preferably at least 15.degree. of crank angle of a nominal
crank shaft driving the piston.
15. A method of controlling the temperature of gas in a gas
compressor during compression, the gas compressor comprising a
compression chamber to contain gas to be compressed, a compression
piston to compress the gas by movement of the compression piston in
the compression chamber, a pressurised source of liquid and a duct
arranged to feed liquid from the pressurised source to the
compression chamber, the method comprising the steps of spraying
liquid into the compression chamber by allowing the source to
accelerate liquid through the duct and controlling the rate of
acceleration of the mass flow into the compression chamber during
compression by sizing the duct to define an inertia of liquid
therein such that liquid is injected into the compression chamber
when the pressure in the compression chamber is greater than the
pressure of the source.
16. A method according to claim 15, wherein the inertia is such
that the period during which liquid is injected while the pressure
in the compression chamber is greater than the pressure of the
source is at least 5.degree., preferably at least 10.degree. and
more preferably at least 15.degree. of crank angle of a nominal
crank shaft driving the piston.
Description
[0001] This invention relates to gas compressors and in particular
to reciprocating gas compressors in which liquid is sprayed into
the compression chamber to control the gas temperature during
compression.
[0002] The concept of spraying liquid into a compression cylinder
as a means for absorbing the heat of compression is described in a
number of publications and referred to in the art as "wet
compression". The technique involves spraying liquid into the
cylinder through a nozzle which divides the liquid into a mist of
fine droplets. The droplets travel through the gas space and
eventually impinge on the cylinder surfaces. While in the gas
space, the droplets provide a heat sink which is in intimate
contact with the gas being compressed and which has a large surface
area allowing heat to be drawn efficiently from the gas, thereby
limiting the rise in gas temperature and reducing the work required
for compression.
[0003] The spray liquid may be driven through the nozzle or nozzles
by a mechanical, positive displacement pump, as, disclosed for
example in GB 722524, FR 903471 and WO 98/16741. On the other hand,
the spray liquid may be driven by the compressed gas from the
compressor, as disclosed in DE 357858. The compressor described in
this document includes an accumulator which temporarily stores
compressed gas and spray liquid drawn from the compression
cylinder. The lower part of the accumulator is connected to a heat
exchanger for cooling the spray liquid, and a reservoir containing
cooled liquid below the heat exchanger is connected to a spray
orifice in the compression cylinder by a duct. The liquid spray is
driven and controlled by the pressure in the accumulator without
any active control mechanism. Liquid is sprayed into the cylinder
during both the air intake and compression strokes and ceases when
the pressure in the compression cylinder reaches the pressure in
the accumulator.
[0004] While compressed gas from the compressor offers an
attractive driving means for injecting liquid into the compression
cylinder by removing the need for additional mechanical pumps, a
drawback of this method is the diminished control it provides over
the flow rate into the cylinder compared with the control provided
by a positive displacement pump. For example, the pressure in the
accumulator is limited to the final pressure in the compression
cylinder. Also, the pressure available to drive the spray liquid is
the difference between the pressure in the accumulator and the
pressure in the cylinder and this difference diminishes over the
compression stroke and falls off rapidly to zero as the cylinder
pressure approaches its final value. Thus, the flow rate, which
depends on the pressure difference across the spray nozzle also
decreases over the compression stroke and falls off rapidly as the
cylinder pressure approaches its final value. However, most of the
heat is evolved towards the end of compression, and therefore the
flow rate needs to increase during compression to meet the
increased cooling demand. Although this requirement can be met by
using a positive displacement pump, since it allows the injection
pressure and flow rate to be controlled independently of the
cylinder pressure, this solution involves additional equipment and
a power source.
[0005] In diesel engines, liquid injection systems are used to
inject controlled quantities of liquid fuel into the combustion
cylinder at a point in the cycle when the combustion air in the
cylinder is at high pressure. Fuel injection is commonly driven by
a com-operated mechanical pump to achieve the high injection
pressures required. Another method is to use "common rail
injection" in which the fuel is pumped into a common pressurised
reservoir and then admitted by small very fast acting valves into
each cylinder in turn. More recently, it has been proposed to drive
the fuel injection using the "water hammer" effect. C. Stan and E.
Hilliger "Pilot Injection System for Gas Engines using
Electronically Controlled Ram Tuned Diesel Injection", CIMAC
Congress 1998, Copenhagen, pages 1429 to 1438 describes a pilot
fuel injection system for a gas engine comprising a closed loop
circuit through which diesel fuel is circulated, a fuel injector
whose input is tapped into the circuit and an electronically
controlled valve placed in the circuit, downstream from the
connection to the injector, for controlling the fuel flow. To
operate the injector, the valve is opened and fuel is driven
through the circuit by a pressure source which includes a fuel pump
and pressure accumulator. In this phase, no fuel is injected since
the pressure of fuel in the circuit is below the injector threshold
pressure. To operate the injector, the valve is rapidly closed,
abruptly blocking the flow through the circuit and generating a
pressure wave and rapid increase in pressure sufficient to drive
fuel through the injector. The peak pressure depends, inter alia,
on the final velocity of the liquid in the circuit just prior to
closing the valve which can be varied by varying the period over
which the valve is open. As the pressure wave duration is constant
and independent of pressure amplitude, the system allows the
injection volume to be conveniently controlled by changing the
pressure in the hydraulic circuit. However, this system does not
allow the injection duration nor the injection profile to be easily
controlled and requires the operation of a mechanical pump.
[0006] According to the present invention there is provided a gas
compressor comprising a compression chamber to contain gas to be
compressed, a compression piston to compress the gas by movement of
the compression piston in the compression chamber, valve means for
allowing compressed gas to be drawn from the compression chamber,
an atomiser for spraying liquid into the compression chamber to
absorb heat from the gas during compression, a pressurised source
of liquid and a duct arranged to feed liquid from the pressurised
source to the atomiser, wherein the source is arranged to
accelerate the liquid through the atomiser into the compression
chamber and the dimensions of the duct are sized to define an
inertia of the liquid therein to control the rate of acceleration
of the mass flow through the atomiser during compression such that
the flow rate of liquid into the compression chamber is
substantially reduced when the pressure difference between the
source and the compression chamber is high and substantially
enhanced when the pressure difference between the source and the
compression chamber is low.
[0007] Advantageously, the present invention uses the inertia of
liquid in the flow path between the pressurised source and spray
atomiser to control the flow rate of liquid into the compression
chamber. For example, the inertia can be defined to control the
rate of acceleration and deceleration of liquid through the
atomiser during compression. Thus, in the initial stages of
compression, when the pressure difference between the pressurised
source and compression cylinder is large but the rate of heat
evolution is low, the inertia of the liquid can be defined to
moderate or limit the rate of acceleration of liquid through the
atomiser and limit the flow rate of liquid into the compression
chamber. Advantageously, this allows the flow rate to be controlled
so that only a sufficient quantity of liquid is present in the gas
to provide the modest heat absorption capacity required in the
early stages of compression. Advantageously, the inertia provides a
means of reducing the mass flow rate without consuming and
dissipating energy, the energy being stored in the liquid and being
used to continue to drive liquid through the atomiser against the
increase in pressure in the compression chamber. The inertia also
provides a means of delaying the highest flow rates through the
atomiser until the pressure in the compression chamber and
therefore the heat of compression is rising rapidly as the pressure
approaches its final value. The increase in flow rate is limited to
some extent as the flow resistance increases sharply at high flow
rates. Furthermore, the inertia also controls the rate of
deceleration of liquid through the atomiser. For example, a high
inertia causes the flow to continue at a relatively high rate near
the end of compression in spite of the diminished pressure
difference between the pressurised source and the compression
chamber.
[0008] As mentioned above, a degree of control over the
acceleration and deceleration of the liquid is provided by defining
the inertia. However, in order to enhance the flow control, a valve
is preferably provided to control the timing of the flow through
the atomiser. Thus, the valve can be used to prevent water from
being injected into the compression chamber during the air intake
period. During this time, no heat is generated and the injection of
water which does not perform efficient cooling may simply increase
the parasitic power loss as work has to be done by the compressor
piston in order to expel this water from the cylinder at high
pressure. With the valve, the start of the injection can be delayed
until the temperature of the air under compression has risen to a
point where efficient heat transfer can take place. In compressors
running at high speed, it may be advantageous to begin the
injection of water during the air intake period, in order that the
water is distributed within the cylinder when the temperature
starts to rise. In this case also, the valve can be used to control
the start of injection in order to maximise the beneficial cooling
effect with minimum usage of water. The valve may also be used to
terminate the flow quickly at the end of the compression period.
This may cause some of the kinetic energy of the flow in the pipe
to be lost. However, it will prevent the injection from extending
unnecessarily into the air discharge period.
[0009] The inertia of liquid in a duct is proportional to its
length and inversely proportional to the cross-sectional area of
the flow passage defined thereby and therefore the inertia can be
conveniently controlled according to those parameters.
[0010] Preferably, the dimensions of the duct are sized to define
the inertia such that the mass flow increases through the atomiser
in a period of at least 30.degree., more preferably at least
45.degree. and most preferably at least 60.degree.of crank angle of
a nominal crank shaft driving the piston.
[0011] Advantageously, the present invention allows the flow rate
to be controlled during compression without the need for a positive
displacement pump and without the need for a pressure source of
greater pressure than the compressed gas drawn from the compression
chamber. Conveniently, the pressurised source may comprise a
reservoir of liquid pressurised by compressed gas and conveniently,
the compressed gas may be supplied by the compression chamber.
[0012] Preferably the reservoir is an accumulator, and the
compressor further comprises a separator upstream of the
accumulator which receives gas and liquid from the compression
chamber, means to feed the liquid in the separator to the
accumulator, and a cooler to cool the liquid being fed from the
separator to the accumulator. The use of the accumulator isolates
the pumping system from the upstream circuit. Without it the
inertia of the upstream pipework would affect the pumping through
the atomisers. Further, by providing a separate accumulator and
separator, space is available for the cooler to be provided to
extract from the liquid the heat absorbed in the compression
chamber.
[0013] Preferably, a further duct is provided to feed liquid from
the reservoir to the atomiser. The dimensions of the further duct
are preferably sized to define an inertia of liquid therein
different to the inertia defined by the dimensions of the first
duct. A further valve is preferably provided for controlling the
flow of liquid through the further duct. With this arrangement,
under the control of the valve and the further valve, the duct with
the lower inertia can be used to contribute a higher mass through
the atomiser sooner and the duct with the higher inertia can be
used to contribute a higher flow rate through the atomiser
later.
[0014] The gas compressor may comprise a further atomiser for
spraying liquid into the compression chamber, the flow through the
atomiser being controlled by a further valve. This arrangement may
be particularly advantageous when the atomisers are arranged to
direct liquid towards different positions within the compression
chamber displaced in the direction of movement of the piston. For
example, one atomiser may be arranged to spray liquid into the
volume adjacent the end of the compression chamber and the other
atomiser may be arranged to spray liquid into the volume spaced
from the end of the compression chamber. The use of ducts having
different inertias feeding different atomisers allows the flow rate
of liquid through the atomisers and into the compression chamber to
be controlled individually so that the flow rate through individual
atomisers may be different during the period of compression.
[0015] As described above, the duct and the further duct may be
provided entirely separately from one another. However, they are
preferably connected by a branch duct. This allows the flow through
one of the atomisers to be shut off, whereupon the flow to the
shut-off atomiser is directed through the branch duct to the other
atomiser. Thus when one of the atomisers is covered by the
compressor piston, this atomiser can be shut off without losing the
benefit of the kinetic energy of the water upstream of the branch
duct. If a branch duct is used, a third valve is preferably
provided in one of the ducts upstream of the location where the
branch duct joins the duct, while the valve and further valve are
downstream of the branch duct. Normally the third valve would be
placed in the upstream duct with the lower inertia. In this case,
the third valve takes over the function of controlling the point at
which flow begins in the duct with the low inertia.
[0016] Also according to the present invention there is provided a
method of controlling the temperature of gas in a gas compressor
during compression, the gas compressor comprising a compression
chamber to contain gas to be compressed, a compression piston to
compress the gas by movement of the compression piston in the
compression chamber, a pressurised source of liquid and a duct
arranged to feed liquid from the pressurised source to the
compression chamber, the method comprising the steps of spraying
liquid into the compression chamber by allowing the source to
accelerate liquid through the duct and controlling the rate of
acceleration of the mass flow into the compression chamber during
compression by sizing the duct to define an inertia of liquid
therein such that the flow rate of liquid into the compression
chamber is substantially reduced when the pressure difference
between the source and the compression chamber is high and
substantially enhanced when the pressure difference between the
source and the compression chamber is low.
[0017] Alternatively, the invention can be defined in its broadest
sense as a gas compressor comprising a compression chamber to
contain gas to be compressed, a compression piston to compress the
gas by movement of the compression piston in the compression
chamber, valve means for allowing compressed gas to be drawn from
the compression chamber, an atomiser for spraying liquid into the
compression chamber to absorb heat from the gas during compression,
a pressurised source of liquid and a duct arranged to feed liquid
from the pressurised source to the atomiser, wherein the source is
arranged to accelerate the liquid through the atomiser into the
compression chamber and the dimensions of the duct are sized to
define an inertia of the liquid therein to control the rate of
acceleration of the mass flow through the atomiser during
compression such that liquid can be injected into the compression
chamber when the pressure in the compression chamber is greater
than the pressure of the source.
[0018] By allowing injection of liquid against a negative pressure
gradient between the source and the compression chamber, the
invention allows the liquid to be injected towards the latter part
of the piston stroke when the compression chamber pressure and
hence the heat of compression is at its highest.
[0019] This aspect of the invention also extends to a method of
controlling the temperature of gas in a gas compressor during
compression, the gas compressor comprising a compression chamber to
contain gas to be compressed, a compression piston to compress the
gas by movement of the compression piston in the compression
chamber, a pressurised source of liquid and a duct arranged to feed
liquid from the pressurised source to the compression chamber, the
method comprising the steps of spraying liquid into the compression
chamber by allowing the source to accelerate liquid through the
duct and controlling the rate of acceleration of the mass flow into
the compression chamber during compression by sizing the duct to
define an inertia of liquid therein such that liquid is injected
into the compression chamber when the pressure in the compression
chamber is greater than the pressure of the source.
[0020] Preferably, the inertia is such that the period during which
liquid is injected while the pressure in the compression chamber is
greater than the pressure of the source is at least 5.degree.,
preferably at least 10.degree. and more preferably at least
15.degree. of crank angle of a nominal crankshaft driving the
piston.
[0021] The preferred features defined above in relation to the
first aspect of the invention will apply equally to the alternative
definition of the invention given above.
[0022] Examples of embodiments of the present invention will now be
described with reference to the drawings, in which:
[0023] FIG. 1 shows a graph of the variation of liquid flow rate as
a function of crank angle with negligible liquid inertia;
[0024] FIG. 2 shows a graph of the variation of pressure difference
across a spray nozzle as a function of crank angle with negligible
liquid inertia;
[0025] FIG. 3 shows a graph of the variation of flow rate as a
function of crank angle with increasing liquid inertia;
[0026] FIG. 4 shows a graph of the pressure across a spray nozzle
as a function of crank angle with increasing liquid inertia;
[0027] FIG. 5 shows the variation of pressure across a spray nozzle
as a function of crank angle for two different values of duct
resistance defining similar liquid inertias;
[0028] FIG. 6 shows an embodiment of the present invention;
[0029] FIG. 7A shows a graph of a variation of pressure as a
function of crank angle for the embodiment shown in FIG. 6; and
[0030] FIG. 7B shows a graph of the mass flow rate as a function of
crank angle for the embodiment shown in FIG. 6.
[0031] FIGS. 1 and 2 illustrate, respectively, an example of the
flow rate through and pressure across a spray nozzle as a function
of crank angle where the source of liquid is at the exhaust
pressure of the compressor and the inertia of liquid in the flow
path between the pressurised source and the spray nozzle is
negligible. The compression piston is at bottom dead center at
-180.degree. and at top dead center at 0.degree. crank angle and in
this example, the discharge valve opens at -30.degree. crank angle.
As shown in these Figures, the maximum liquid injection rate occurs
at the start of the compression process, i.e. at -180.degree. crank
angle as the pressure in the compression chamber rises during
compression. The compressor reaches the desired pressure at about
30.degree. before top dead center and from this point onwards, the
pressure in the compression chamber is the same as pressurised
liquid source. Therefore, the differential pressure between the
pressurised source and the compression chamber and hence the flow
rate fall to zero and remain zero from that point to top dead
center.
[0032] However, this type of injection profile may not be the
optimum for ensuring near-isothermal compression. In a
reciprocating compressor, the majority of compression work is done
towards the end of the compression stroke, when the pressure is
rising rapidly and is accompanied by a rapid increase in the rate
at which the heat of compression is evolved. Therefore, it is
important that the cooling capacity of spray liquid and hence the
mass of cool liquid in transit through the compression chamber is
sufficient to absorb the rapid increase in heat energy as the
pressure in the cylinder approaches its final value. However, in
FIG. 1, the liquid injection rate rapidly falls off towards the end
of compression. Also, in FIG. 1, the flow rate is high over the
initial stages of compression when the compression work and hence
heat evolution is relatively low and the large cooling capacity
provided by the spray droplets is not required. Although a valve
could be used to restrict the flow rate during the early stages of
compression, the valve would cause energy to be dissipated which is
undesirable. Also the use of a valve in this way would not help the
injection rate later in the process, when the pressure difference
between the source and cylinder becomes much smaller.
[0033] FIGS. 3 and 4 illustrate, respectively, examples of flow
rate through and pressure across a spray nozzle as a function of
crank angle where the flow path between the pressurised source and
the nozzle has an inertia which is sufficient to modify the flow
rate profile. Again, the pressurised source is at the same pressure
as the final pressure in the compression chamber and the discharge
valve is opened at about 30.degree. before top dead center. In FIG.
3, the inertia of the liquid in the flow path increases between
curves A to C and in FIG. 4, the pressure difference across the
spray nozzle in the latter stages of the compression increases from
curve D to F. Referring to FIGS. 3 and 4, as the gas pressure acts
on the liquid, the liquid in the flow path is accelerated at a rate
which decreases with increasing inertia. Therefore, as the flow
begins, in the early stages of compression, the inertia causes an
effective pressure drop by reducing the flow and differential
pressure across the spray nozzle. Thus, the inertia can be used to
control the rate of increase of liquid flow through the atomiser
during compression.
[0034] Resistive losses due to the ductwork and spray atomiser will
also produce a pressure drop between the pressurised source and
compression chamber. This will cause the liquid to begin to
decelerate before the compression chamber reaches the source
pressure, and this effect is shown in FIG. 3 by the decreasing flow
rate before the discharge valve is opened at 30% before top dead
center. However, during deceleration, caused by the increase in
pressure in the compression chamber, the liquid inertia reduces the
rate of deceleration, producing an effective increase in
differential pressure across the atomiser, relative to the case
without inertia, thereby prolonging and sustaining a relatively
high mass flow through the atomiser as the pressure in the
compression chamber reaches its final value and sustaining a mass
flow into the compression chamber after the gas pressure has
reached its final value.
[0035] The inertia, I, of liquid in a duct of length L and internal
cross-sectional area A is given by the expression I=L/A.
[0036] The pressure drop along the duct due to the inertia is:
P.sub.2-P.sub.1=I dW/dt m.sup.-1
[0037] where W is the mass flow rate in kg/s and P is pressure
(N/m.sup.2)
[0038] Thus, when the liquid is being accelerated by the
pressurised source, the inertia causes an effective pressure drop
and reduction in acceleration of the mass flow and when the liquid
is being decelerated by the rising pressure in the compression
chamber and frictional losses in the ductwork, the inertia causes
the liquid to react against the decelerative forces causing an
effective pressure increase.
[0039] Changing the duct length or bore will also affect the duct
resistance, so whereas the inertia effect will change the shape of
the injection profile, the resistance affects both the magnitude of
the atomiser pressure drop and the flow rate.
[0040] FIG. 5 shows two cases where the liquid in the connecting
duct between the pressurised source and compression chamber has a
similar inertia in both cases, but the lower curve H has a much
higher flow resistance and therefore lower flow and lower atomiser
pressure drop than curve G.
[0041] The resistance is proportional to the duct length and
inversely proportional to the fifth power of the duct diameter and
so varies in a different way to the inertia. Thus, by choosing
appropriate duct cross-sectional areas and lengths, and by choice
of valve opening and closing times the desired profile can be
obtained. The flow rate can also be reduced by throttling the
pressure from the water reservoir which provides a way of raising
the flow resistance without affecting the inertia
characteristic.
[0042] Throttling the flow or having a high flow resistance leads
to losses in the liquid supply system. At times this may be
acceptable, since the liquid supply reservoir will be at a high
pressure and it may not be desirable to use this full pressure
across the atomisers all of the time. However, towards the end of
the stroke, the losses should be minimised and the inertia effect
used to ensure a good pressure drop and flow rate. To minimise the
losses, the cross-sectional area of the duct should not be too
small, and consequently to ensure a high inertia, the ductwork and
duct length will need to be relatively long.
[0043] An example of the present invention will now be described
with reference to FIG.6. A cylinder 1 defines a compression chamber
2 in which gas is compressed by a reciprocating compression piston
3. The gas to be compressed is fed into the compression chamber 2
through gas inlet line 4 under the control of gas inlet valve
5.
[0044] During the compression operation, water is injected as a
spray of fine droplets through upper 6 and lower 7 elevationally
spaced annular nozzle manifolds. The arrangement of nozzles is
described in more detail in WO 96/16741.
[0045] The cool compressed gas together with the injected water are
drawn out of the compression chamber 2 through gas outlet line 8
under the control of gas outlet valve 9. The compressed gas and
water are fed to a separator 10 where the gas and water are
separated. The gas is removed through line 11 for subsequent
heating and expansion. The water is fed to a heat exchanger 12
where it is cooled. The recovered heat may be used in another part
of the cycle, or for space heating, or may be dumped. The water
leaving the heat exchanger 12 is fed to an accumulator 13. The
accumulator 13, which contains a compressible gas, can deliver a
high flow rate for a short duration into the inertial pipe system.
The accumulator acts as the upstream limit of the pipework which
provides the inertial flow.
[0046] The liquid is fed to the nozzle manifolds 6, 7 through a
system of ducts and valves as described below. A first duct 14
connects the accumulator 13 to a second duct 15 which leads to the
upper nozzle manifold 6. Flow through the first duct 14 is
controlled by first control valve 16 and flow through the second
duct 15 is controlled by second control valve 17 downstream of the
first control valve. The first duct 14 has a relatively low
inertia, such that, when the valve 16 opens, the flow in the first
duct 14 accelerates rapidly. A third duct 18 consists of a large U
bend having a relatively high inertia leading from the accumulator
13 and a fourth duct 19 of shorter length than the third duct 18
connects the third duct 18 to the lower nozzle manifold 7. The flow
through the fourth duct 19 is controlled by a third control valve
20. No valve is required in the third duct 18 since the flow in
this duct begins earlier than the flow in the first duct 14 and can
therefore be controlled by second 17 and third 20 control valves. A
branch duct 21 connects the first 14 and second 15 ducts to the
third 18 and fourth 19 ducts.
[0047] The first to third control valves 16, 17, 20 are fast acting
valves such as hydraulic spool valves. On the other hand, the
valves 5, 9 on the cylinder 1 are typically poppet valves.
[0048] Before the injection begins, all three control valves 16,
17, 20 are closed. Injection is normally initiated shortly after
the start of compression by opening both second 17 and third 20
control valves. This is shown beginning about at -165.degree. in
FIGS. 7A and 7B. The control valves 17, 20 can be opened at the
same time, or one can be opened before the other. For example,
third control valve 20 could be opened sooner since this would
allow more time for the water to penetrate to the lower part of the
cylinder 1. Thus, water flows to both nozzle manifolds 6, 7 during
the early part of compression, with the flow rate rising as shown
in FIG. 7B. Since the first control valve 16 is still closed at
this time, the flow builds up relatively slowly because of the high
inertia of third duct 18.
[0049] As the piston 3 rises up the cylinder 1, the work rate and
therefore the heat input to the cylinder 1 increase. It therefore
becomes desirable to increase the water injection rate more
rapidly, and therefore the first control valve 16 is opened at
about -130.degree.. Since the first duct 14 upstream of the first
control valve 16 has a low inertia, the flow rate to both of the
nozzle manifolds increases rapidly. FIG. 7B illustrates the
resulting increase in flow rate, particularly in the upper manifold
6, to which there is a more direct path from the accumulator 13.
The boost to the flow provided by the opening of the first control
valve 16 helps to overcome the increasing flow resistance from the
nozzles and the associated pipework.
[0050] Near the end of the compression stage, when the work and
heat input is near its maximum, the piston 1 begins to obstruct or
completely block flow from the lower nozzle manifold 7. This is
typically at around -50.degree. crank angle. Further injection of
water from the lower manifold 7 is wasted since most of it simply
hits the piston and does not contribute significantly to the
cooling. At this time, the third control valve 20 is closed to
block the flow to the lower manifold 7. This also has the effect of
forcing all of the water to flow to the upper nozzle manifold 6
providing a boost to this pressure as shown in FIG. 7A. As is also
apparent from this figure, the pressure in the lower manifold 7
beyond this stage will simply follow the pressure in the cylinder.
Since there is still significant kinetic energy in the upstream
ducts 14, 15, 18, 19, 21 at this time, the closure of the third
control valve 20 provides a further boost to the flow rate to the
nozzles of the upper manifold 6 which are arranged to direct the
flow into a relatively narrow space at the top of the cylinder 1.
The pressure rise caused by the closing of the third control valve
20 is sufficient to overcome the rapidly rising cylinder pressure
and thus the high rate to the upper manifold is extended until the
end of the compression at -30.degree. as shown in FIG. 7A. The
pressure is boosted to such an extent that for crank angles between
40.degree. and 20.degree. before top dead center, water is injected
even though the cylinder pressure is above the source pressure.
[0051] When the discharge valve 9 opens, it is desirable to shut
off the water flow as quickly as possible without causing
unacceptable pressures due to water hammer. This may be done by
closing the second control valve 16 over an appropriate time scale.
Finally, first control valve 16 is also closed in readiness for the
next cycle.
[0052] It should be noted that this embodiment uses the compressed
air in the separator 10 as the source of pressure to drive the
injection of the liquid. However, it would be equally possible to
derive the pressure for the pumping system from a different source,
such as a conventional pump. The function of the accumulator 12, on
the other hand, is to smooth out the pressure fluctuations, so that
the inertia of all upstream pipework between the separator and the
accumulator does not affect the behaviour of the inertial pumping
system. It is likely that an accumulator would still be required in
such an alternative system using a conventional pump.
[0053] As an alternative to the cooler 12 and accumulator 13, it is
possible to provide a cooler in the bottom of the tank 10. This
could take the form of a coiled pipe through which a coolant liquid
flows. As well as eliminating the need for the cooler 12, this
arrangement could also remove the need for the accumulator 13. In
this case, the separator 10 would have to be close enough to the
cylinder 1 that the pipework could still be arranged to provide the
required inertia.
[0054] Although FIG. 6 indicates that the inertia required by the
third duct 18 could be achieved by a single U-bend, this might not
be sufficient in some applications. Alternative means of achieving
the desired pipe inertia within a confined space include having a
number of coils in the pipe, or connecting a number of U-bends and
straight lengths together in series to form a serpentine
configuration similar to that used in some heat exchanger
designs.
[0055] Additional nozzle manifolds could be provided with
additional valves to control the flow through them. Also additional
ducts could be connected to the accumulator 13. These additional
ducts may have different inertias and may also have valves to
control the flow in those ducts to start at different times. A
single branch duct could connect all the ducts leading from the
accumulator 13 to all the ducts leading to the various atomisers.
Alternatively, connections could be made selectively, for example
involving more than one branch duct.
* * * * *