Chemical and Petroleum Engineering, Vol. 37, Nos. 9–10, 2001
CRYOGENIC ENGINEERING,
PRODUCTION AND USE OF INDUSTRIAL GASES
PISTON EXPANDER-COMPRESSOR UNIT HAVING
SELF-ACTING GAS DISTRIBUTION SYSTEMS
A. D. Vanyashov, V. S. Kalekin,
and S. V. Kovalenko
UDC 621.575.56.001.5
The major area where gas-expansion machines (expanders) are employed is air separation and gas liquefaction
plants [1]. Operation of such expanders at a low initial air pressure (as low as 0.8 MPa) to produce moderate temperatures
is inefficient due to impairment of specific energy parameters. This is associated primarily with forced gas distribution systems used in piston (reciprocating) expanders.
In a forced valve drive (external and internal), there is a large number of highly passive (sluggish) moving parts,
because of which the rotation speed of piston expanders is low (as low as 500 rpm with an external drive and 1000 rpm with
an internal drive) [1].
One way of refining designs of piston expanding machines is to replace forced gas distribution systems by self-acting valves [2–4].
In the piston expander cylinder, the gas may be made to move by a concurrent flow or by a nonconcurrent flow
scheme: the concurrent flow scheme provides for a normally open inlet valve and an exit port at the end of the piston stroke
(at the lower dead spot); the nonconcurrent flow scheme provides for a normally open inlet and a closed outlet valve.
Self-acting valves, by virtue of their low inertia, allow one to raise the rotation speed of the expander crankshaft to
the level of the rotation speed of modern high-speed piston compressors, which makes it possible to combine a compressor
and an expander into an expander-compressor unit (ECU) with a common crannkshaft.
Expander-compressor units designed to obtain moderate temperatures and operating at a low pressure can be built
on unified compressor bases with a rated piston force of the order of 2.5–16 kN.
In recent years, several designs have been developed and investigations have been carried out into the processes that
occur in piston expanders having self-acting valves [5].
Let us examine a piston ECU having an expander operating in accordance with the concurrent flow [6] and nonconcurrent flow [7] gas distribution schemes.
The experimental investigations of the ECU were carried out on a bench built on the basis of the vertical two-stage
two-set ship compressor 20K1 which has one-way operating differential pistons: cylinder diameters 0.1 and 0.035 m respectively, piston stroke 0.1 m, rated rotation speed 500 rpm, air suction (intake) pressure of the compressor atmospheric, and
compression ratio 4–8.
The compressor 20K1 was converted to an ECU as follows.
The valve heads of the cylinders with a diameter 0.035 m of the second stage of the compressor were replaced by
valve heads with self-acting valves that ensure gas distribution of the function of the expander that operates at a low initial
pressure (as low as 1 MPa). In the lower part of one of the cylinders (diameter 0.035 m), a circular slit was made on the
perimeter for the exit of the expanded cooled air.
Omsk State Technical University. Translated from Khimicheskoe i Neftegazovoe Mashinostroenie, No. 9, pp. 28–31,
September, 2001.
474
0009-2355/01/0910-0474$25.00 ©2001 Plenum Publishing Corporation
The experimental bench offered the possibility for implementing three gas distribution schemes in the expander
stage.
Concurrent flow scheme: entry of compressed air through a normally open self-acting inlet valve, exit of expanded air through an exit slit in the expander cylinder.
Nonconcurrent flow scheme: entry of compressed air through a normally open self-acting inlet valve, exit of
expanded air through a self-acting outlet valve.
Combined scheme: entry of compressed air through a normally open self-acting inlet valve, exit of expanded air
through an exit slit in the expander cylinder and a self-acting outlet valve.
For the first scheme, four versions of a ring-type self-acting inlet valve have been designed and made. For the second
and third schemes, a valve system containing plate-type self-acting inlet and outlet valves has been designed and made. The
experiments were conducted by turn on one of the sets of the expander stage, the second set being turned off during this time.
The parameters of the self-acting inlet valve for the concurrent flow gas distribution scheme are: ring diameter
0.035 and 0.037 m, ring width 0,012 and 0.013 m, ring thickness 0.001, 0.0015, and 0.002 m, number of holes in the seat 7,
number of springs 1, diameter of the hole in the seat 0.0045 m, dead volume of the valve Vd.v = 6397 mm3, relative dead volume a = 0.063, outer diameter of springs 0.018 m, and spring force Cspr = 823, 1029, and 1370 N/m.
The parameters of the self-acting outlet valve for nonconcurrent flow gas distribution scheme are: plate diameter
0.016 m, plate thickness 0.001 m, number of holes in the seat 1, number of springs 1, diameter of the hole in the seat 0.01 m,
outer diameter of springs 0.005 m, and spring force Cspr = 263, 823, 1100, 1500, 2100, and 2800 N/m.
The parameters of the self-acting inlet valve for nonconcurrent flow gas distribution scheme are: plate diameter
0.019 m, plate thickness 0.001 m, number of holes in the seat 1, number of springs 1, diameter of the hole in the seat 0.015 m,
outer diameter of springs 0.005 m, and spring force Cspr = 70, 100, 263, 823, 1100, 1500, and 2100 N/m.
Since the last two valves are installed in the same valve head, their total dead volume Vd.v = 8493 mm3 and relative
dead volume a = 0.088.
The designs of the four versions of the inlet valve (for the concurrent flow scheme) and two versions of the outlet
valve (for the nonconcurrent flow and combined schemes) make it possible to regulate the height of lift of the cut-off element
hmax from 0 to 0.005 m by shifting the lift limiter relative to the seat on a thread having a fine pitch. The designs of the inlet
valve for the nonconcurrent flow and combined schemes make it possible to change the height of lift of the cut-off element
at a constant spring tension as well as to change the spring tension at a constant height of lift of the cut-off element.
Maintenance of a constant height of lift of the cut-off element under changing spring tension does not lead to a rise
in hydraulic losses in the filling process and ensures permissible speed of seating of the cut-off element on the seat. Maintenance of a constant spring tension under changing height of lift of the cut-off element makes it possible to ensure the required
spring resilience force in case the degree of cylinder filling cut-off is regulated under noncalculated operation conditions.
On the experimental bench, the following were measured: the rapidly changing pressure in the expander and compressor cylinders and the outlet pipeline of the expander, the rapidly changing temperature in the expander cylinder, and the
diagrams of movement of the cut-off elements of the self-acting valves of the expander stage. Provision was also made for
recording of the position of the piston in the upper and lower dead spots (UDS and LDS). Also measured were the external
parameters: the compressor suction (intake) pressure and the expander inlet pressure, the temperature at the compressor suction and at the expander inlet and outlet, the rate of air flow into the compressor, and the electric power consumed by the ECU.
Results of Experimental Studies of the ECU by the Concurrent Flow Gas Distribution Scheme. The cut-off element lift height hmax, spring force Cspr, and dead volume affect the working process of the expander stage.
In Fig. 1 is shown the experimental dependence of rapidly changing pressure in the inlet pipeline and in the expander
and compressor cylinders on the shaft turning angle ϕ for one of the conditions (hmax = 0.75 mm and Cspr = 840 N/m) and
in Fig. 2, the diagram of movement of the cut-off element of the valve.
The thermodynamic efficiency of the expander stage depends a great deal on the relative piston stroke at the moment
of closing of the inlet valve c2 (angle ϕ2), i. e., on the degree of cut-off of expander cylinder filling, as well as on the relative
piston stroke at the moment of opening of the inlet valve c6 (angle ϕ6).
475
Fig. 1. Experimental diagrams of pressures in ECU having a self-acting
valve (Cspr = 840 N/m, hmax = 0.75 mm, and a = 0.09): 1) in inlet pipeline,
2) in compressor cylinder, and 3) in expander cylinder.
Fig. 2. Experimental diagram of movement of the cut-off element of the
valve (Cspr = 840 N/m, hmax = 0.75 mm, and a = 0.09).
It has been established that to each spring force value corresponds a fixed maximum valve lift height at which valve
closing takes place at ϕ2 = 90° (c2 = 0.55). To these conditions correspond the minimum air temperature at the expander outlet Te = 270 K and the maximum refrigeration power of the unit Q = 180 W.
By approximation of the experimental data, we obtained equations that express the relationship hmax = ƒ(Cspr) at
c2 = 0.55 and by experiments, the dependence of c2 on hmax at variable Cspr, from which it follows that the parameter c2
increases with increase of hmax at a given spring force and with increase of spring force at constant hmax.
As the valve opening angle ϕ6 approaches the UDS, air flow through the expander and refrigeration power of the
unit increase, and the work for compression of the residual air diminishes. Based on the experimental data, it was concluded that the valve parameters hmax and Cspr have little effect on the parameter c6. The dead volume of the expander stage exercises a great influence on the angle ϕ6. With a 4.2-fold rise of a (from 0.063 to 0.267), the valve opening angle ϕ6 rises by
20° (from 300 to 320°).
Because of the peculiar nature of operation of the ECU, the pressure at the expander stage inlet (compressor discharge pressure) is not constant and is dependent on the parameter c2.
In the course of the experiment, the pressure at the expander stage inlet, upon adherence to the valve closing condition (ϕ2 = 90°), remained more or less constant (0.65 MPa). If the valve remains open for a long time (ϕ2 ≥ 140°), the mass
air flow through the expander increases and the pressure at the expander inlet drops to 0.4–0.5 MPa. If the valve closes prematurely (ϕ2 ≤ 50°), the mass air flow through the expander becomes less than the compressor delivery, which causes the
compressor discharge pressure and the pressure in the inlet pipeline to rise to 0.8 MPa. If hmax is too low (for a given spring
476
Fig. 3. Dependence of indicated power of compressor Nc (a) and of expander Ne (b) on degree of cut-off of
cylinder filling c2.
force), the valve closes almost instantly in the first few cycles after start of the ECU. The pressure in the pipeline reaches
0.9–1 MPa and during the back stroke of the piston the valve may not open at all.
The energy parameters of the ECU (power consumed by the compressor and returned by the expander) also depend
a great deal on the parameter c2. The indicated power of the compressor stage Nind.c depends linearly on c2 (Fig. 3a). The
decrease in power consumption by the compressor with increase of c2 occurs due to a drop in the compressor discharge pressure at high filling degrees. The change in the indicated power of the expander stage Nind.e is of a parabolic pattern and has
an optimum (maximum value) at c2 = 0.5–0.55 (Fig. 3b). For the expander stage, the maximum returnable power, determined
experimentally, was 140 W. The drop in Nind.e at lower c2 values occurred due to the short filling process, even though the
working cycle occurs at high initial pressures. At higher c2 values the indicated power also diminishes, which is attributable
to an increase in the proportion of the work of reverse compression.
The ratio of the indicated powers of the expander and the compressor (Nind.e /Nind.c) is proportional to the fraction
of the power returned to the ECU shaft by the expander stage.
Results of Experimental Studies of Nonconcurrent Flow and Combined Gas Distribution Schemes. Of great
importance (besides the parameters c2 and c6) is the degree of cut-off of the discharge c5 (relative piston stroke at the moment
of closure of the outlet valve), which should be increased in order to attain maximum power and refrigerating ability of the
expander stage. The influence of the design of the inlet and outlet valves on the parameters c2 and c5 and on the expander
efficiency was studied.
Three versions of operation of the outlet valve were examined:
a) the directions of action of the spring force and the gas pressure force coincide;
b) the directions of action of the spring force and the gas pressure force are opposite;
c) without spring.
Convoluted indicator diagrams of the expander stage for the above-referred cases of outlet valve operation are shown
in Fig. 4.
Experimental Results:
• use of normally closed outlet valve (the spring presses the cut-off element to the seat) is ineffectual. For outlet valve
opening, it is necessary to expand the air to a pressure below the atmospheric, which causes the degree of cylinder filling to drop.
In the case of back stroke of the piston, the direction of action of the force of the outlet valve spring coincides with the direction
of the gas pressure force, which facilitates early closure of the valve and reduces the duration of the ejection process (Fig. 4a);
• use of normally open outlet valve (without a spring) raised expander efficiency, but the duration of the filling and
ejection process remained unchanged (Fig. 4b);
477
Fig. 4. Indicator diagrams of expander stage: a) outlet valve with a spring that presses the cut-off element
to the seat, b) outlet valve without a spring, c) outlet valve with a spring that detaches the cut-off element from
the seat, and d) combined gas distribution system.
• use of normally open outlet valve (the spring detaches the cut-off element from the seat) made it possible to obtain
near-optimum (corresponding to operation at a low pressure) angles of closing and opening of both valves. For opening of
the outlet valve, it is not necessary to expand the air to the atmospheric pressure (which raises the degree of cylinder filling).
During the back stroke of the piston, the force of the outlet valve spring opposes the gas force, which facilitates more delayed
closing of the valve and increases the duration of the ejection process. However, the outlet valve closing angle could not be
increased to the optimum (90°) because in that case the pressure drop at the end of expansion is not enough for opening of
the outlet valve because of large air volume in the cylinder. Use of a more rigid spring on the outlet valve leads to a situation where, during the back stroke of the piston, it either does not close or closes very late and the pressure at the end of the
reverse compression process is inadequate for opening of the outlet valve (Fig. 4c).
For the nonconcurrent flow gas distribution scheme, based on experimental data, c2 ≤ 0.3 is recommended. Above
this value, this scheme is inefficient. The degrees of intake and discharge cut-off depend on both the relative dead volume
and the rigidity of the outlet valve spring.
Application of the combined gas distribution scheme at the outlet ensures a fairly high degree of filling (c2 ≤ 0.5)
over a fairly long ejection time (c5 = 0.5). In this series of experiments, use was made of only normally open position of the
outlet valve. Opening of the exit slit by the piston on account of increasing pressure differential facilitates opening of the
outlet valve. In this case, it is possible to make use on the inlet valve of springs of fairly high rigidity required to ensure
c5 = 0.5. The indicator diagram for this case is shown in Fig. 4d.
An analysis of the operation of the three possible gas distribution schemes showed that in the combined gas distribution scheme, at c2 = 0.3, the indicated power and refrigeration output are as much as 30% higher than those in the non-
478
concurrent flow scheme and as much as 40% higher than those in the concurrent flow scheme; at c2 = 0.5, they are as much
as 20% higher than those in the concurrent flow scheme.
The results of these investigations experimentally confirmed for the first time the possibility of operation of
expanders having self-acting valves as components of the ECU. The designed self-acting valves are capable of operating at
reduced pressures at the expander inlet. However, the temperature at the expander outlet (270 K) is not low enough, which
is explicable by the presence of heat bridges between the compressor and expander cylinders. In order to reduce the influence on the temperature at the expander outlet, the sets of compressor and expander stages must be spread out among themselves. Therefore, use of general-service and sucker-rod type of compressor bases, one set of which will be expander, is most
promising for the creation of ECU.
REFERENCES
1.
2.
3.
4.
5.
6.
7.
A. M. Arkharov, K. S. Butkevich, I. K. Butkevich, and A. Z. Mirkin, Cryogenic Piston Expanders [in Russian],
A. M. Arkharov (ed.), Mashinostroenie, Moscow (1974), p. 240.
Russian Federation Patent No. 2029911, ICI F 25 B 1/02, Piston Expander.
I. K. Prilutskii and A. I. Prilutskii, Calculation and Designing of Piston Compressors and Expanders on Standardized Bases. Text Book for Institutes of Higher Education [in Russian], SPGAKhiPT, St. Petersburg (1995), p. 193.
Yu. I. Molodova, “Analysis of operation of piston expansion machines,” Kompress. Tekh. Avtomat., No. 18–19,
37–41 (1998).
A. D. Vanyashov, V. S. Kalekin, and A. N. Kabakov, “Expander-compressor unit having self-acting valves,” Kryogen. Kholod. Oborud., Part 2, 216–224 (1999).
Piston Expander. Certificate for Uutility Model No. 11312, ICI F 25 B 1/02.
Piston Expanding Machines. Certificate for Uutility Model No. 13060, ICI F 01 L 9/02 and F 01 B 25/02.
479