The HallScrew compressor for refrigeration and heat pump duties

C. Y. Chan , G. G. H a s e l d e n and G. H u n d y

Le compresseur HallScrew utiliser pour le froid et la pompe de chaleur On d~crit la construction du compresseur Ha//Screw qui s'appuie sur la d#couverte du compresseur monovis par Zimmern et I'on in(lique ses principaux avantages. On utilise un programme d'ordinateur de

g#om#trie de base pour rechercher les avantages de modifications des proportions de la machine, y compris le nombre de rainures et de dents des rotors. On rend compte de la performance pour une s#rie de machines normales fonctionnant au R 122 et au R 717, y compris I'influence du refoulement, de /'injection de Iiquide et du fonctionnement de I'#conomiseur. On donn¢ qu#lques indications sur le fonctionnement de la pompe de chaleur utilisant le R 114.

The construction of the HallScrew compressor, based on the Zimmern single- screw invention, is described and its basic advantages are outlined. A computer model of the basic geometry is used to explore the advantages of changes in the proportions of the machine, including the numbers of flutes

The HallScrew compressor is a version of the single- screw compressor invented by Bernard Zimmern. 1 An alternative version, called the Monoscrew, having the same basic geometry was described by J. van Male, 2 and a comparison of this machine wi th the SRM twin-screw (Lysholm) machine was presented by A. Lundberg and R. Glanvall. 3

The fundamental virtues of the single-screw design are:

One, high throughput for the given rotor size because each flute in the main rotor is used twice per revolution.

Two, very l ight bearing loads on the main rotor because both the radial and axial pressure loads are balanced, leaving little more than the weight of the rotor to be carried. This fact also makes the maintenance of very small clearances much easier.

Three, the gate-rotors do no work, but have only a sealing function. Hence the work to be transmitted to them is very small indeed, therefore rubbing loads wi th in the machine are very light.

and teeth on the rotors. Performance is reported for a range of standard machines operating on R 22 and R 717 including the effects of unloading, liquid injection and economiser operation. Some data are given for heat pump conditions using R 114.

internal construction is presented in Fig. 2. The special features include:

The one-piece casing. By very careful attention to the form of the gate-rotor slots in the cylindrical casing of the compressor, and to the design of the bearings, it was possible to insert the gate-rotors into the main rotor after it had been installed. This insertion is diff icult because the gate-rotors must be threaded into the main rotor using a twist ing action - they cannot be fed straight in. The advantages are many-fold, the casing becomes a far stronger pressure vessel, a long gasketed joint is avoided wi th its influence on important clearances, and the

Description of the Hal lScrew design

A photograph of a HalIScrew compressor is shown in Fig. 1 and a cut-away diagram showing its

CYC and GGH are from the Department of Chemical Engineering. University of Leeds. Leeds, UK. GH is from HalI-Thermotank Products Ltd, Home Gardens, Dartford, Kent DA1 1EP, UK. Paper received 13 May 1981.

Fig. 1 Photograph of a HallScrew refrigeration compressor

Fig. I Photo du compresseur frigorifique Hal/Screw

0140-7007/81 /050275-0652.00 Volume 4 Num~ro 5 Septembre 1981 @1981 IPC Business Press Ltd and IIR 275

openings, and the pressure loss highlighted by Lundberg and Glanvall 3 is almost eliminated. Churning losses are also significantly reduced. Incidentally, the use of the plastic gate-rotors enables the sealing line to be brought very close to the rotor surface so that tile blow-hole mentioned by these authors virtually disappears also.

The form of the unloading slides, previously described, 4 enabled the throughput to be reduced to below 20% of full capacity whilst also affording some reduction of pressure ratio over a substantial part of this range, as is needed by the normal operational requirement.

The main purpose of this paper is to examine the basic geometry of the single-screw globoid compressor and to present operating data for a range of machines in refrigeration and heat pump duties.

IJ

Fig. 2 Cutaway diagram showing constructional features of the HallScrew compressor

Fig. 2 Vue en coupe des caractdri~tiques de construction du compresseur He#Screw

gate-rotors can be inspected in position and can be removed or replaced without disturbing the rest of the machine.

The main rotor is made of cast iron whilst the sealing portion of the gate-rotors is made from a composite plastic supported on a steel backing. The use of cast iron gives excellent dimensional stability, an ideal surface finish and a much lower thermal conductivity than the bronze or aluminium alloy previously used. The latter point is important at the discharge end of the rotor where steep temperature gradients can arise. The use of plastic teeth on the gate-rotors allows (with the correct choice of filled plastic) a low rubbing friction combination with cast iron so that some dry running can occur without damage, and very long running periods can be achieved provided some liquid coolant is present. Thus the plastic/cast iron combination provides freedom to operate with far less oil present or with oil that has become greatly diluted with refrigerant. The plastic also has a thermal conductivity which is much lower than that of metals.

The compressor casing is designed so that the cylindrical inner surface is relieved in all areas except those required for the actual compression. By this modification the low pressure gas is able to enter the flutes radially and not just through the end

Basic geometry

All the single-screw compressors so far marketed employ a 6-flute main rotor meshing with 11 -teeth gate-rotors. The outside diameters of the main and gate-rotors have been about equal and the degree of interpenetration has been about 45% of the rotor radius. All these dimensions are independently capable of variation, leading to variations of swept volume, potential compression ratio, lengths of leakage paths and other machining and operating factors. If a general method of modelling the geometry can be developed then it will be possible to analyse how optimal is the 6-11 geometry, and what further improvements are possible.

The form of the flutes in the main rotor is best defined by following the profile of a gate-rotor tooth as it moves progressively through it, both rotors moving at the appropriate angular velocities. Fig. 3 shows the main variables, D being the diameter of the main rotor, R the radius of the star and S the centre-line spacing. The gate-rotor teeth are parallel

Suction side

,%,o

Discharge

G

:4- 4 j /

[ ] AT, engaged tooth area

Fig. 3 Basic dimensions used in defining the rotors and their inter-relationships

Fig. 3 Donn~es dimensionnelles utilisdes pour d~finir les rotors et leurs relations entre eux

276 International Journal of Refrigeration

o4 I

Fig. 4 rotor

Mode 5

The modes of engagement of the star teeth with the main

Fig. 4 Modes d'engagement des dents en #teile sur le rotor principal

sided and of width B, such that for the given value of S the teeth do not totally penetrate the main rotor - otherwise the flute walls would become yanishingly thin. The following assumptions are made:

One, the clearance between the flute walls and the edges of the teeth are ignored, ie, the flute and tooth widths are taken as equal. Similarly the flute depth is defined by the length of the tooth.

Two, for small rotational increments of the main rotor the globoid groove is assumed to follow an annular path around the rotor axis, and the exposed tooth area A T, and its moment about the rotor axis u, may be assumed constant.

If the rotor revolves through a small angle &e causing the star to rotate by &8 then

&t7 = 1-~&~#

for the 6 flute, 11 teeth geometry. Also the increment of swept volume is given by

AVs =AT X u x1--~6 tlt7

and the total flute volume

disengagement V T = ~./t i V s

cut-off

It remains therefore to define AT and u as functions of 8, and to specify the limits of 8.

In considering A T it was found convenient to separate three modes of engagement of a tooth, as

shown in Fig. 4. The first was when both parallel side edges of the tooth are engaged in the main rotor, and this mode occupies most of the motion. The second mode occurs when only one straight edge is engaged but the centre-line of the tooth is inside the main rotor boundary, whilst the third is when the centre-line is outside the boundary. In the case of the first mode the tooth surface is divided into three simpler shapes which are easily defined. In the second and third modes the areas can be approximated by triangles. Once the engaged tooth area is defined it is relatively easy to obtain its centre of gravity and hence the distance u.

Fig. 5 illustrates the problem of defining the values of 8 at the commencement of compression (cut-off) and disengagement. It will be appreciated that the volume of the flute is symmetrical about the centre line '00', therefore the total stroke can be compounded from two parts, one covering values of 8>90 ° and the other for 8<90 ° . Cut-off occurs when the tooth comes completely within the profile of the main rotor, and hence is influenced by the amount of truncation. In fact there is no point in having cut-off occurring before the corresponding tooth in the other gate-rotor has disengaged from the opposite end of the same flute. This consideration determines the amount of truncation, and hence the cut-off angle. The value of 8 at the termination of discharge is more straightforward, and is determined directly by the rotor and tooth dimensions.

A computer program was written which incorporated these factors. The various dimensions were normalised with respect to the main rotor diameter so that any throughput could in theory be

A , f / f

0 - - . - - ~ - - - ~ 0

\ /

Fig. 5 The limits of tooth engagement at cut-off and at discharge

Fig. 5 Limites d'engagement de la dent lots de I'arr~t et au refoulement

Volume 4 Number 5 September 1981 277

0.04

An interesting product of this computation was the effect on volumetric displacement of the degree of penetration of the gate-rotor into the main rotor. Fig. 6 shows the two segments of the flute volume on either side of 0=90" and V T is the total. It is seen that displacement rises to a fairly flat maximum when the gate-rotor has penetrated by about 40% of the main rotor radius. For deeper penetration the decreasing tooth width outweighs the advantage of greater tooth length. The degree of penetration may also be limited by machining requirements since the flank angle of the main rotor flutes which can be achieved must be related to the form of the cutting tool and its motion.

Table 1 shows the change of swept flute volume for the same main rotor diameter which could be achieved, at least theoretically, by using different numbers of flutes in the main rotor (from 4 to 8), and different numbers of teeth on the gate-rotors (from 5 to 1 6 in appropriate combinations). In each case the degree of penetration is optimised. The 6-11 combination is used as standard.

It would appear that very significant throughput increases could be achieved by going to main rotors

Table 1. Effect of different combinations of main rotors and gate-rotors on maximum swept volume

Tableau 1. Influence de diffdrentes combinaisons des rotors principaux et des rotors-vannes sur le volume balayd maximal

Configuration % Deviation of swept volume No of flutes No of teeth from 6-11

4 5 +50.7

6 9 +22.4

4 6 +24.4 6 10 +10.9 8 14 +10.4

4 7 +4.0 6 11 Standard

4 8 -11.4 6 12 -10.0 8 16 -8.0

Table 2. Range of machines currently in production

Tableau 2. Gamme des machines en cours de construction

Model Rotor diameter, Swept volume at mm 3000 rpm m 3 s -~

HS24 245 0.24

HS28 280 0.35

HS31 310 0.47

HS35 350 0.685

vT 0,03

0.02

,T

0.01

handled by scaling any machine with respect to its main rotor diameter.

0 I I I I i 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2

Degree of Nr~trotion

Fig. 6 Effect of penetration of the gate-rotor teeth into the main rotor on volumetr ic displacement

Fig. 6 Influence sur le d6bit volum~rique de la p6n~ration des dents en #toile clans le rotor principal

with four flutes rather than six and by using the minimum number of teeth. This achievement would be limited in practice both by machining limitations and by the compression ratios that can be handled. Other problems arise when there is a common factor between the numbers of flutes and teeth.

P e r f o r m a n c e

Range of machines. The range of machines currently in production is shown in Table 2. With each machine there is provision for inbuilt volumetric ratios of 2.2-4.9 corresponding to pressure ratios of about 3.0-7.5.

With the introduction of the single piece casing, the opportunity was taken to increase gas flow areas at the inlet, discharge port and for part-load gas by- passing. These, together with the other design changes, have resulted in improvements in performance throughout the entire operating range, as compared to the performance achieved with the prototype units previously reported. 3

As production machines have become available for detailed testing, so the opportunity has been taken to obtain data enabling the performance to be mapped.

Performance for refrigeration duties. Figs 7 and 8 show the performance curves thus obtained for the full range of compressors on both R 22 and ammonia (R 717).

In common with all positive displacement rotary compressors, the degree of inbuilt compression is determined by the built-in volume ratio, VR. This is the ratio of the flute volume at the start of compression to that when the delivery port opens.

The volumetric efficiency curve in each case is a composite one and bridges between the different in- built volume ratios, being relatively insensitive to this factor. It is apparent in Fig. 7 for R 22 that the efficiency curves have fairly flat maxima, as with all screw compressors, but that significant gains are won by selecting the right volumetric ratio. Also in

278 Revue Internationale du Froid

each range the effect of size of machine is apparent. The larger machines gain because throughput and power are proportional essentially to the cube of the rotor diameter whilst the losses arising from leakage. heat transfer and friction are proportional to the linear or square power of the diameter. Thus both volumetric and isentropic efficiency improve with scale in this size range.

With ammonia it is seen that there are significant regions of operation in which the isentropic efficiency is above 80%, especially for the larger machines, and efficiencies above 75% are general up to compression ratios of about 6.0.

Part-load performance of two machines is demonstrated in Fig. 9. With the largest machine at

t00

90 Volumetric.

5 70 p

4.9 I ~ t m ~ c

60 ~ 2 3 4 5 6 7 8 9 I0 I I 12 13 14 15

Pres=ure ratio

Fig. 7 HallScrew compressor efficiency when operating wi th R 22. The numbers on the isentropic efficiency curves are the volumetric ratios. For each set of four curves the upper curve is for the largest machine (350 mm rotor) the others being for 310, 280, 240 mm diameter main rotors, respectively

Fig. 7 Rendement du compresseur HallScrew fonctionnant au R 22.Les nombres sur les courbes de rendement isentropique repr~sentent les taux volum~riques. Pour cl~aque ensemble de quatre courbes la courbe supdrieure est destinde ~ la plus grosse machine (rotor de 350 mm), les autres ~ celle du rotor respectivemem de 310, 280 et 240 mm de diam~re

ioo

90

8(:

7C

6C

2 3 4 5 6 7 8 9 I0 I I 12 13 14 15

Pmlmure ratio

Fig. 8 HallScrew compressor efficiency when operating wi th ammonia - otherwise as in Fig. 7

Fig. 8 Rendement du compresseur HallScrew fonctionnant rammoniac. Pour le reste se reporter ~ la Fig. 7

80

~60

UJ

40

20 0

Fig. 9

~ . 9

H S ~ i~l~um rotio 5 .7

I I I I 20 40 60 80

Capacity, %

Part-load isentropic efficiency, refrigerant R 717

Rendement isentrol2ique ~ charge partielle. R 717

IOO

a pressure ratio of 5.7 it is seen that the isentropic efficiency is better than 70% down to 65% of full capacity, and even at 40% of full capacity the efficiency is only slightly below 6•%. The smaller machine has an efficiency of better than 60% down to 50% of full-load. These tests were conducted at constant operating conditions. Advantage can then be taken of the proportionately larger heat transfer areas available at part-load on a refrigeration plant. This can result in reduction of discharge pressure with additional consequent power reduction.

Liquid refrigerant injection is an attractive alternative to oil cooling because it avoids the need for a costly water cooled oil cooler and its associated maintenance problems. Also. where cooling water is not available, this mode of operation becomes essential.

Tests have been made with liquid injection directly into the compression chambers with both R 22 and R 717. The object of these tests has been to establish reliability and effect on performance. The presence of a very low viscosity liquid in the compression chambers could possibly reduce the lubricant film at the gate-rotor contact region to a point where wear starts to occur. However, an HS28 compressor has operated with excessive amounts of liquid R 22 (2-5 m 3 h -1) for a period of over 2000 h with no apparent effect. Typically the volumetric efficiency is slightly reduced at high pressure ratios when liquid injection is used to control the discharge temperature at 75°C, this reduction is due to increased leakage of refrigerant, either in solution in the oil or in liquid form. This effect can be minimised by controlling at a higher discharge temperature.

Provision is made on the compressors for two intermediate suction ports. These ports are identically positioned on each side of the compressor, one for each compression process. They are exposed to the compression chambers for part of the compression cycle and this enables an additional charge of gas to be handled by the compressor over and above the normal suction.

Volume 4 Num(~ro 5 Septembre 1 981 279

I I CondenlNlr

Fig. 10

Fig. 10

I Compressor

: l

Thermostatic I expansion valve

J Sub-cooler

Economiser circuit diagram

Diagramme du circuit dconomiseur

Economiser

Suction

Evaporator

t IOO

ae

LLI

90

80 ̧

70:

60

50

40 I I I I 1 2 3 4 5 6 7

Pressure rol"io

Fig. 11 Performance test data for HallScrew HS24 compressor operating on R 114. Condensing temperature 100-120°C, discharge superheat 15-20°C

Fig. 11 R~suitats d'essai de performance du compresseur HalIScrew HS24 fonctionnant au R 114. Tempdrature de condensation; 100 ~ 120°C, surchauffe au refoulement: 15 ~ 20°C

The usual method of application of this economiser facility isto provide the additional gas at the intermediate pressure by evaporating liquid refrigerant in the heat exchanger in order to sub- cool the remaining liquid refrigerant (Fig. 10).

For the HS28 compressor, the evaporator duty gain with R 22 at -40/35°C (evaporator temperature, condenser temperature) and at optimum vapour injection is 31% and at conditions of -50/35°C it is 38%. The respective power increase at these conditions was measured to be 7% and 5% respectively and so it can be seen that a substantial

improvement in refrigerating performance is possible with this method.

Heat pump application of the HallScrew compressor

Elevation of the condensing pressure of a R 12 heat pump to provide a condensing temperature of up to 75°C, presents no problems with a screw type machine. It has the ability to operate at high pressure ratios in a single stage if required to do so, and then discharge temperature can be controlled by oil cooling or direct injection of liquid refrigerant.

When higher temperatures are being considered R 114 is used as refrigerant. The advantages of the HallScrew arise mainly as a result of the low demands made on oil viscosity. By contrast, the twin-screw compressor depends on the maintenance of an oil film to transmit power from one rotor to another. Also the main bearings and thrust faces are heavily loaded and require good lubrication. The test results shown in Fig. 11 are for the HS24 compressor operating on R 114.

Extensive tests have been made with liquid R 22 injection into the compression chambers, and these have shown that provided the oil film at the bearings is maintained, and that the shaft seal is adequately lubricated, this mode of operation presents no mechanical problems. However as the discharge temperature approaches saturation temperature, an increasing amount of refrigerant is retained in solution in the oil in the separator. This is particularly true with R 114 which has a relatively high solubility in most oils. With an oil injected machine therefore, difficulties can arise with oil separation if discharge superheat is allowed to fall too low.

Conclusions

The reported results show the rate at which the development of the single-screw compressor is enabling it to compete with, and in some cases overhaul, the performance of other compressors. In particular it possesses a very high degree of flexibility which enables it to cover a wide range of duties, including heat pumps, and to take advantage of liquid injection and economiser circuits.

References

1 Zimrnern, B., Patel, G. C. Design and operating characteristics of the Zimmern single-screw compressor, Purdue Compressor Technology Conference (1972)

2 van Malo, J. Monoscrew - a newly developed refrigeration compressor. Int J Refr I (1978) 242-8

3 Lundberg, A., Glanvall , R. A comparison of SRM and GIoboid type screw compressors. Int J Refr 2 (1979) 221-232

4 Hundy, G. F. Part load operating and testing of the single- screw compressor with refrigerant vapours, I Mech E Conference: Design and Operation of Industrial Compressors, Strathclyde (20-22 March 1978)

280 International Journal of Refrigeration